Stirling cycle machine

ABSTRACT

A Stirling cycle machine. The machine includes at least one rocking drive mechanism which includes: a rocking beam having a rocker pivot, at least one cylinder and at least one piston. The piston is housed within a respective cylinder and is capable of substantially linearly reciprocating within the respective cylinder. Also, the drive mechanism includes at least one coupling assembly having a proximal end and a distal end. The proximal end is connected to the piston and the distal end is connected to the rocking beam by an end pivot. The linear motion of the piston is converted to rotary motion of the rocking beam. Also, a crankcase housing the rocking beam and housing a first portion of the coupling assembly is included. A crankshaft coupled to the rocking beam by way of a connecting rod is also included. The rotary motion of the rocking beam is transferred to the crankshaft. The machine also includes a working space housing the at least one cylinder, the at least one piston and a second portion of the coupling assembly. A seal is included for sealing the workspace from the crankcase.

CROSS REFERENCE TO RELATED APPLICATIONS

The present application is a Continuation of U.S. patent applicationSer. No. 12/105,854, filed Apr. 18, 2008 and entitled Stirling CycleMachine, now U.S. Pat. No. 8,474,256, issued Jul. 2, 2013, which claimspriority from U.S. Provisional Patent Application No. 60/925,818, filedApr. 23, 2007 and entitled Four Cylinder Stirling Engine; and U.S.Provisional Patent Application No. 60/925,814, filed Apr. 23, 2007 andentitled Rocking Beam Drive, all of which are hereby incorporated hereinby reference in their entireties.

TECHNICAL FIELD

The present invention relates to machines and more particularly, to aStirling cycle machine and components thereof.

BACKGROUND INFORMATION

Many machines, such as internal combustion engines, external combustionengines, compressors, and other reciprocating machines, employ anarrangement of pistons and drive mechanisms to convert the linear motionof a reciprocating piston to rotary motion. In most applications, thepistons are housed in a cylinder. A common problem encountered with suchmachines is that of friction generated by a sliding piston resultingfrom misalignment of the piston in the cylinder and lateral forcesexerted on the piston by linkage of the piston to a rotating crankshaft.These increased side loads increase engine noise, increase piston wear,and decrease the efficiency and life of the engine. Additionally,because of the side loads, the drive requires more power to overcomethese frictional forces, thus reducing the efficiency of the machine.

Improvements have been made on drive mechanisms in an attempt to reducethese side loads, however, many of the improvements have resulted inheavier and bulkier machines.

Accordingly, there is a need for practical machines with minimal sideloads on pistons.

SUMMARY

In accordance with one aspect of the present invention, a rocking beamdrive mechanism for a machine is disclosed. The drive mechanism includesa rocking beam having a rocker pivot, at least one cylinder and at leastone piston. The piston is housed within a respective cylinder. Thepiston is capable of substantially linearly reciprocating within therespective cylinder. Also, the drive mechanism includes at least onecoupling assembly having a proximal end and a distal end. The proximalend is connected to the piston and the distal end is connected to therocking beam by an end pivot. The linear motion of the piston isconverted to rotary motion of the rocking beam.

Some embodiments of this aspect of the present invention include one ormore of the following: where the rocking beam is coupled to a crankshaftby way of a connecting rod. In this embodiment, the rotary motion of therocking beam is transferred to the crankshaft. Also, where the cylindermay further include a closed end and an open end. The open end furtherincludes a linear bearing connected to the cylinder. The linear bearingincludes an opening to accommodate the coupling assembly. Also, wherethe coupling assembly further includes a piston rod and a link rod. Thepiston rod and link rod are coupled together by a coupling means. Thecoupling means is located beneath the linear bearing. Also, where thedrive mechanism also includes a seal, where the seal is sealablyconnected to the piston rod. Also, where the seal is a rollingdiaphragm. Also, in some embodiments, the coupling means is a flexiblejoint. In some embodiments, the coupling means is a roller bearing. Insome embodiments, the coupling means is a hinge. In some embodiments,the coupling means is a flexure. In some embodiments, the coupling meansis a journal bearing joint.

In accordance with another aspect of the present invention, a Stirlingcycle machine is disclosed. The machine includes at least one rockingdrive mechanism where the rocking drive mechanism includes: a rockingbeam having a rocker pivot, at least one cylinder and at least onepiston. The piston is housed within a respective cylinder. The piston iscapable of substantially linearly reciprocating within the respectivecylinder. Also, the drive mechanism includes at least one couplingassembly having a proximal end and a distal end. The proximal end isconnected to the piston and the distal end is connected to the rockingbeam by an end pivot. The linear motion of the piston is converted torotary motion of the rocking beam. Also, a crankcase housing the rockingbeam and housing a first portion of the coupling assembly is included. Acrankshaft coupled to the rocking beam by way of a connecting rod isalso included. The rotary motion of the rocking beam is transferred tothe crankshaft. The machine also includes a working space housing the atleast one cylinder, the at least one piston and a second portion of thecoupling assembly. A seal is included for sealing the workspace from thecrankcase.

Some embodiments of this aspect of the present invention include one ormore of the following: where the seal is a rolling diaphragm. Also, thecylinder may further include a closed end and an open end. The open endfurther includes a linear bearing connected to the cylinder. The linearbearing includes an opening to accommodate the coupling assembly. Also,where the coupling assembly further includes a piston rod and a linkrod. The piston rod and link rod are coupled together by a couplingmeans. The coupling means may be located beneath the linear bearing.Also, the machine may also include a lubricating fluid pump in thecrankcase. In some embodiments, the lubricating fluid pump is amechanical lubricating fluid pump driven by a pump drive assembly, thepump drive assembly being connected to and driven by the crankshaft. Insome embodiments, the lubricating fluid pump is an electric lubricatingfluid pump. The machine may also include a motor connected to thecrankshaft. The machine may also include a generator connected to thecrankshaft.

In accordance with another aspect of the present invention, a Stirlingcycle machine is disclosed. The machine includes at least two rockingdrive mechanisms. The rocking drive mechanisms each include a rockingbeam having a rocker pivot, two cylinders, and two pistons. The pistonseach housed within a respective cylinder. The pistons are capable ofsubstantially linearly reciprocating within the respective cylinder.Also, the drive mechanisms include two coupling assemblies having aproximal end and a distal end, the proximal end being connected to thepiston and the distal end being connected to the rocking beam by an endpivot. The linear motion of the piston is converted to rotary motion ofthe rocking beam. The machine also includes a crankcase housing therocking beam and housing a first portion of the coupling assemblies.Also, a crankshaft coupled to the rocking beam by way of a connectingrod. The rotary motion of the rocking beam is transferred to thecrankshaft. The machine also includes a lubricating fluid pump in thecrankcase for pumping lubricating fluid to lubricate the crankshaft andthe rocking beam and the first portion of the coupling assemblies. Also,a working space housing the cylinders, the pistons and the secondportion of the coupling assemblies. A rolling diaphragm for sealing theworkspace from the crankcase is also included.

Some embodiments of this aspect of the present invention include one ormore of the following: where the cylinder may further include a closedend and an open end. The open end further includes a linear bearingconnected to the cylinder. The linear bearing includes an opening toaccommodate the coupling assembly. Also, where the coupling assemblyfurther includes a piston rod and a link rod. The piston rod and linkrod are coupled together by a coupling means. The coupling means may belocated beneath the linear bearing. Also, where the coupling means is aflexible joint. In some embodiments, also disclosed is where thecoupling means is a roller bearing.

These aspects of the invention are not meant to be exclusive and otherfeatures, aspects, and advantages of the present invention will bereadily apparent to those of ordinary skill in the art when read inconjunction with the appended claims and accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

These and other features and advantages of the present invention will bebetter understood by reading the following detailed description, takentogether with the drawings wherein:

FIGS. 1A-1E depict the principle of operation of a prior art Stirlingcycle machine;

FIG. 2 shows a view of a rocking beam drive in accordance with oneembodiment;

FIG. 3 shows a view of a rocking beam drive in accordance with oneembodiment;

FIG. 4 shows a view of an engine in accordance with one embodiment;

FIGS. 5A-5D depicts various views of a rocking beam drive in accordancewith one embodiment;

FIG. 6 shows a bearing style rod connector in accordance with oneembodiment;

FIGS. 7A-7B show a flexure in accordance with one embodiment;

FIG. 8 shows a four cylinder double rocking beam drive arrangement inaccordance with one embodiment;

FIG. 9 shows a cross section of a crankshaft in accordance with oneembodiment;

FIG. 10A shows a view of an engine in accordance with one embodiment;

FIG. 10B shows a crankshaft coupling in accordance with one embodiment;

FIG. 10C shows a view of a sleeve rotor in accordance with oneembodiment;

FIG. 10D shows a view of a crankshaft in accordance with one embodiment;

FIG. 10E is a cross section of the sleeve rotor and spline shaft inaccordance with one embodiment;

FIG. 10F is a cross section of the crankshaft and the spline shaft inaccordance with one embodiment;

FIG. 10G are various views a sleeve rotor, crankshaft and spline shaftin accordance with one embodiment;

FIG. 11 shows the operation of pistons of an engine in accordance withone embodiment;

FIG. 12A shows an unwrapped schematic view of a working space andcylinders in accordance with one embodiment;

FIG. 12B shows a schematic view of a cylinder, heater head, andregenerator in accordance with one embodiment;

FIG. 12C shows a view of a cylinder head in accordance with oneembodiment;

FIG. 13A shows a view of a rolling diaphragm, along with supporting topseal piston and bottom seal piston, in accordance with one embodiment;

FIG. 13B shows an exploded view of a rocking beam driven engine inaccordance with one embodiment;

FIG. 13C shows a view of a cylinder, heater head, regenerator, androlling diaphragm, in accordance with one embodiment;

FIGS. 13D-13E show various views of a rolling diaphragm duringoperation, in accordance with one embodiment;

FIG. 13F shows an unwrapped schematic view of a working space andcylinders in accordance with one embodiment;

FIG. 13G shows a view of an external combustion engine in accordancewith one;

FIGS. 14A-14E show views of various embodiments of a rolling diaphragm;

FIG. 15A shows a view of a metal bellows and accompanying piston rod andpistons in accordance with one embodiment;

FIGS. 15B-15D show views of metal bellows diaphragms, in accordance withone embodiment;

FIGS. 15E-15G show a view of metal bellows in accordance with variousembodiments;

FIG. 15H shows a schematic of a rolling diaphragm identifying variousload regions;

FIG. 15I shows a schematic of the rolling diaphragm identifying theconvolution region;

FIG. 16 shows a view of a piston and piston seal in accordance with oneembodiment;

FIG. 17 shows a view of a piston rod and piston rod seal in accordancewith one embodiment;

FIG. 18A shows a view of a piston seal backing ring in accordance withone embodiment;

FIG. 18B shows a pressure diagram for a backing ring in accordance withone embodiment;

FIGS. 18C and 18D show a piston seal in accordance with one embodiment;

FIGS. 18E and 18F show a piston rod seal in accordance with oneembodiment;

FIG. 19A shows a view of a piston seal backing ring in accordance withone embodiment;

FIG. 19B shows a pressure diagram for a piston seal backing ring inaccordance with one embodiment;

FIG. 20A shows a view of a piston rod seal backing ring in accordancewith one embodiment;

FIG. 20B shows a pressure diagram for a piston rod seal backing ring inaccordance with one embodiment;

FIG. 21 shows views of a piston guide ring in accordance with oneembodiment;

FIG. 22 shows an unwrapped schematic illustration of a working space andcylinders in accordance with one embodiment;

FIG. 23A shows a view of an engine in accordance with one embodiment;

FIG. 23B shows a view of an engine in accordance with one embodiment;

FIG. 24 shows a view of a crankshaft in accordance with one embodiment;

FIGS. 25A-25C show various configurations of pump drives in accordancewith various embodiments;

FIG. 26A show various views of an oil pump in accordance with oneembodiment;

FIG. 26B shows a view of an engine in accordance with one embodiment;

FIG. 26C shows another view of the engine depicted in FIG. 26B;

FIGS. 27A and 27B show views of an engine in accordance with oneembodiment;

FIG. 27C shows a view of a coupling joint in accordance with oneembodiment;

FIG. 27D shows a view of a crankshaft and spline shaft of an engine inaccordance with one embodiment;

FIG. 28 shows a view of a heater exchanger and burner for an engine inaccordance with one embodiment;

FIG. 29 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 30 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 31 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 32 shows a view of heater tubes of a heat exchanger in accordancewith one embodiment;

FIG. 33 shows a view of heater tubes of a heat exchanger in accordancewith one embodiment;

FIG. 34 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 35 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 36 shows a view of a heater head of an engine in accordance withone embodiment;

FIG. 37 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 38 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 39 shows a portion of a cross section of a tube heat exchanger inaccordance with one embodiment;

FIG. 40 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 41 shows a portion of a cross section of a tube heat exchanger inaccordance with one embodiment;

FIG. 42 shows a view of a heater head of an engine in accordance withone embodiment;

FIG. 43A shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 43B shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 44A shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 44B shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 45A shows view of a tube heat exchanger in accordance with oneembodiment;

FIG. 45B shows a view of a tube heat exchanger in accordance with oneembodiment;

FIGS. 46A-46D show various configurations of a tube heat exchanger inaccordance with various embodiments;

FIG. 47 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 48 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIG. 49 shows a view of a heater head of an engine in accordance withone embodiment;

FIG. 50 shows a view of a tube heat exchanger in accordance with oneembodiment;

FIGS. 51A and 51B show views of heat exchangers of an engine inaccordance with various embodiments;

FIGS. 52A-52C show various views of a heat exchanger in accordance withone embodiment;

FIG. 52D shows a view of a heat exchanger in accordance with oneembodiment;

FIGS. 53A and 53B show views of a heat exchanger in accordance with oneembodiment;

FIG. 53C shows a view of a heat exchanger of an engine in accordancewith one embodiment;

FIGS. 53D-53F show views of a heat exchanger of an engine in accordancewith one embodiment;

FIGS. 54A and 54B show views of a heat exchanger of an engine inaccordance with one embodiment;

FIGS. 55A-55D show various views of a heat exchanger in accordance withone embodiment;

FIGS. 56A-56C show various configurations of a heat exchanger inaccordance with various embodiments;

FIGS. 57A and 57B show various diagrams depicting physical properties ofa heat exchanger in accordance with one embodiment;

FIG. 58 shows a view of a heater head in accordance with one embodiment;

FIG. 59 shows a view of a heater head in accordance with one embodiment;

FIGS. 60A and 60B show views of a heater head in accordance with oneembodiment;

FIGS. 61A and 61B show views of a heater head in accordance with oneembodiment;

FIGS. 62A and 62B show views of a heater head in accordance with oneembodiment;

FIG. 62C shows a views of a heater head in accordance with oneembodiment;

FIG. 62D shows a view of a heater head in accordance with oneembodiment;

FIG. 62E shows a view of a heater head in accordance with oneembodiment;

FIGS. 63A and 63B show a regenerator of a Stirling cycle engine inaccordance with one embodiment;

FIGS. 64A-64E show various configurations of a regenerator of a Stirlingcycle engine in accordance with various embodiments;

FIGS. 65A-65G show various views of an engine in accordance with severalembodiments;

FIGS. 66A and 66B show views of a cooler for an engine in accordancewith some embodiments;

FIG. 67A shows a view of a cooler for an engine in accordance with oneembodiment;

FIG. 67A-1 shows an enlarged section view of a section of FIG. 67A;

FIG. 67B shows a view of a cooler for an engine in accordance with oneembodiment;

FIG. 67B-1 shows an enlarged section view of a section of FIG. 67B;

FIG. 67C shows a view of the embodiment of a cooler for an enginedepicted in FIG. 67A;

FIG. 68 shows a view of an intake manifold for an engine in accordancewith one embodiment;

FIGS. 69A and 69B show various views of an intake manifold for an enginein accordance with one embodiment;

FIG. 70 shows a view of a heater head of an engine in accordance withyet another embodiment of the invention;

FIGS. 71A and 71B show views of a burner of an engine in accordance withone embodiment;

FIG. 72 is a gaseous fuel burner coupled to a Stirling cycle engine,where the ejector is a venturi, according to one embodiment;

FIG. 73A is the burner of FIG. 72 showing the air and fuel flow paths;

FIG. 73B is a graphical representation of the pressure across theburner;

FIG. 74 shows a view of a venturi as shown in the burner of FIG. 72;

FIGS. 75A and 75B are embodiments of the venturi in FIG. 72;

FIG. 75C shows a schematic of a multiple fuel system with multiple fuelrestrictions and valves;

FIG. 76 shows a schematic of an embodiment of the burner with automatedfuel control for variable fuel properties;

FIG. 77 shows a schematic of another embodiment of the burner withtemperature sensor and engine speed control loop;

FIG. 78 shows a schematic of yet another embodiment of the burner withtemperature sensor and oxygen sensor control loop;

FIG. 79 shows an alternative embodiment of the ejector wherein the fuelis fed directly into the ejector;

FIG. 80 is a block diagram showing a system for controlling apressurized combustion chamber of an engine according to an embodiment;

FIG. 81 shows a piston pump according to one embodiment;

FIG. 82 shows an alternating current waveform suitable for driving thepiston pump of FIG. 81;

FIG. 83 shows a pulse-width-modulated direct current waveform suitablefor driving the piston pump of FIG. 81, according to one embodiment;

FIG. 84 is schematic diagram of a diaphragm pump according to oneembodiment;

FIG. 85 is a schematic diagram of a center-tapped coil for a diaphragmpump according to one embodiment;

FIGS. 86A and 86B shows pulse-width-modulated direct current waveformssuitable for driving the center-tapped coil of FIG. 85, according tosome embodiments;

FIGS. 87A-87D show embodiments of including a filter between the fuelpump and combustion chamber;

FIG. 88 shows a view of an engine in accordance with one embodiment;

FIGS. 89A-89C show views of a burner for an engine in accordance withvarious embodiments;

FIG. 90 shows a view of an engine with multiple burners in accordancewith yet another embodiment of the invention;

FIGS. 91A and 91B show views of multiple burners for an engine inaccordance with various embodiments;

FIG. 91C shows a view of a tube heater head in accordance with oneembodiment; and

FIG. 91D shows a cross section of the tube heater head depicted in FIG.91C.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Stirling cycle machines, including engines and refrigerators, have along technological heritage, described in detail in Walker, StirlingEngines, Oxford University Press (1980), incorporated herein byreference. The principle underlying the Stirling cycle engine is themechanical realization of the Stirling thermodynamic cycle:isovolumetric heating of a gas within a cylinder, isothermal expansionof the gas (during which work is performed by driving a piston),isovolumetric cooling, and isothermal compression. Additional backgroundregarding aspects of Stirling cycle machines and improvements thereto isdiscussed in Hargreaves, The Phillips Stirling Engine (Elsevier,Amsterdam, 1991), which is herein incorporated by reference.

The principle of operation of a Stirling cycle machine is readilydescribed with reference to FIGS. 1A-1E, wherein identical numerals areused to identify the same or similar parts. Many mechanical layouts ofStirling cycle machines are known in the art, and the particularStirling cycle machine designated generally by numeral 10 is shownmerely for illustrative purposes. In FIGS. 1A to 1D, piston 12 and adisplacer 14 move in phased reciprocating motion within the cylinders 16which, in some embodiments of the Stirling cycle machine, may be asingle cylinder, but in other embodiments, may include greater than asingle cylinder. A working fluid contained within cylinders 16 isconstrained by seals from escaping around piston 12 and displacer 14.The working fluid is chosen for its thermodynamic properties, asdiscussed in the description below, and is typically helium at apressure of several atmospheres, however, any gas, including any inertgas, may be used, including, but not limited to, hydrogen, argon, neon,nitrogen, air and any mixtures thereof. The position of the displacer 14governs whether the working fluid is in contact with the hot interface18 or the cold interface 20, corresponding, respectively, to theinterfaces at which heat is supplied to and extracted from the workingfluid. The supply and extraction of heat is discussed in further detailbelow. The volume of working fluid governed by the position of thepiston 12 is referred to as the compression space 22.

During the first phase of the Stirling cycle, the starting condition ofwhich is depicted in FIG. 1A, the piston 12 compresses the fluid in thecompression space 22. The compression occurs at a substantially constanttemperature because heat is extracted from the fluid to the ambientenvironment. The condition of the Stirling cycle machine 10 aftercompression is depicted in FIG. 1B. During the second phase of thecycle, the displacer 14 moves in the direction of the cold interface 20,with the working fluid displaced from the region of the cold interface20 to the region of the hot interface 18. This phase may be referred toas the transfer phase. At the end of the transfer phase, the fluid is ata higher pressure since the working fluid has been heated at constantvolume. The increased pressure is depicted symbolically in FIG. 1C bythe reading of the pressure gauge 24.

During the third phase (the expansion stroke) of the Stirling cyclemachine, the volume of the compression space 22 increases as heat isdrawn in from outside the Stirling cycle machine 10, thereby convertingheat to work. In practice, heat is provided to the fluid by means of aheater head (not shown) which is discussed in greater detail in thedescription below. At the end of the expansion phase, the compressionspace 22 is full of cold fluid, as depicted in FIG. 1D. During thefourth phase of the Stirling cycle machine 10, fluid is transferred fromthe region of the hot interface 18 to the region of the cold interface20 by motion of the displacer 14 in the opposing sense. At the end ofthis second transfer phase, the fluid fills the compression space 22 andcold interface 20, as depicted in FIG. 1A, and is ready for a repetitionof the compression phase. The Stirling cycle is depicted in a P-V(pressure-volume) diagram as shown in FIG. 1E.

Additionally, on passing from the region of the hot interface 18 to theregion of the cold interface 20. In some embodiments, the fluid may passthrough a regenerator (shown as 408 in FIG. 4). A regenerator is amatrix of material having a large ratio of surface area to volume whichserves to absorb heat from the fluid when it enters from the region ofthe hot interface 18 and to heat the fluid when it passes from theregion of the cold interface 20.

Stirling cycle machines have not generally been used in practicalapplications due to several daunting challenges to their development.These involve practical considerations such as efficiency and lifetime.Accordingly, there is a need for more Stirling cycle machines withminimal side loads on pistons, increased efficiency and lifetime.

The principle of operation of a Stirling cycle machine or Stirlingengine is further discussed in detail in U.S. Pat. No. 6,381,958, issuedMay 7, 2002, to Kamen et al., which is herein incorporated by referencein its entirety.

Rocking Beam Drive

Referring now to FIGS. 2-4, embodiments of a Stirling cycle machine,according to one embodiment, are shown in cross-section. The engineembodiment is designated generally by numeral 300. While the Stirlingcycle machine will be described generally with reference to the Stirlingengine 300 embodiments shown in FIGS. 2-4, it is to be understood thatmany types of machines and engines, including but not limited torefrigerators and compressors may similarly benefit from variousembodiments and improvements which are described herein, including butnot limited to, external combustion engines and internal combustionengines.

FIG. 2 depicts a cross-section of an embodiment of a rocking beam drivemechanism 200 (the term “rocking beam drive” is used synonymously withthe term “rocking beam drive mechanism”) for an engine, such as aStirling engine, having linearly reciprocating pistons 202 and 204housed within cylinders 206 and 208, respectively. The cylinders includelinear bearings 220. Rocking beam drive 200 converts linear motions ofpistons 202 and 204 into the rotary motion of a crankshaft 214. Rockingbeam drive 200 has a rocking beam 216, rocker pivot 218, a firstcoupling assembly 210, and a second coupling assembly 212. Pistons 202and 204 are coupled to rocking beam drive 200, respectively, via firstcoupling assembly 210 and second coupling assembly 212. The rocking beamdrive is coupled to crankshaft 214 via a connecting rod 222.

In some embodiments, the rocking beam and a first portion of thecoupling assembly may be located in a crankcase, while the cylinders,pistons and a second portion of the coupling assembly is located in aworkspace.

In FIG. 4 a crankcase 400 most of the rocking beam drive 200 ispositioned below the cylinder housing 402. Crankcase 400 is a space topermit operation of rocking beam drive 200 having a crankshaft 214,rocking beam 216, linear bearings 220, a connecting rod 222, andcoupling assemblies 210 and 212. Crankcase 400 intersects cylinders 206and 208 transverse to the plane of the axes of pistons 202 and 204.Pistons 202 and 204 reciprocate in respective cylinders 206 and 208, asalso shown in FIG. 2. Cylinders 206 and 208 extend above crankshafthousing 400. Crankshaft 214 is mounted in crankcase 400 below cylinders206 and 208.

FIG. 2 shows one embodiment of rocking beam drive 200. Couplingassemblies 210 and 212 extend from pistons 202 and 204, respectively, toconnect pistons 202 and 204 to rocking beam 216. Coupling assembly 212for piston 204, in some embodiments, may comprise a piston rod 224 and alink rod 226. Coupling assembly 210 for piston 202, in some embodiments,may comprise a piston rod 228 and a link rod 230. Piston 204 operates inthe cylinder 208 vertically and is connected by the coupling assembly212 to the end pivot 232 of the rocking beam 216. The cylinder 208provides guidance for the longitudinal motion of piston 204. The pistonrod 224 of the coupling assembly 212 attached to the lower portion ofpiston 204 is driven axially by its link rod 226 in a substantiallylinear reciprocating path along the axis of the cylinder 208. The distalend of piston rod 224 and the proximate end of link rod 226, in someembodiments, may be jointly hinged via a coupling means 234. Thecoupling means 234, may be any coupling means known in the art,including but not limited to, a flexible joint, roller bearing element,hinge, journal bearing joint (shown as 600 in FIG. 6), and flexure(shown as 700 in FIGS. 7A and 7B). The distal end of the link rod 226may be coupled to one end pivot 232 of rocking beam 216, which ispositioned vertically and perpendicularly under the proximate end of thelink rod 226. A stationary linear bearing 220 may be positioned alongcoupling assembly 212 to further ensure substantially linearlongitudinal motion of the piston rod 224 and thus ensuringsubstantially linear longitudinal motion of the piston 204. In anexemplary embodiment, link rod 226 does not pass through linear bearing220. This ensures, among other things, that piston rod 224 retains asubstantially linear and longitudinal motion.

In the exemplary embodiment, the link rods may be made from aluminum,and the piston rods and connecting rod are made from D2 Tool Steel.Alternatively, the link rods, piston rods, connecting rods, and rockingbeam may be made from 4340 steel. Other materials may be used for thecomponents of the rocking beam drive, including, but not limited to,titanium, aluminum, steel or cast iron. In some embodiments, the fatiguestrength of the material being used is above the actual load experiencedby the components during operation.

Still referring to FIGS. 2-4, piston 202 operates vertically in thecylinder 206 and is connected by the coupling assembly 210 to the endpivot 236 of the rocking beam 216. The cylinder 206 serves, amongstother functions, to provide guidance for longitudinal motion of piston202. The piston rod 228 of the coupling assembly 210 is attached to thelower portion of piston 202 and is driven axially by its link rod 230 ina substantially linear reciprocating path along the axis of the cylinder206. The distal end of the piston rod 228 and the proximate end of thelink rod 230, in some embodiments, is jointly hinged via a couplingmeans 238. The coupling means 238, in various embodiments may include,but are not limited to, a flexure (shown as 700 in FIGS. 7A and 7B,roller bearing element, hinge, journal bearing (shown as 600 in FIG. 6),or coupling means as known in the art. The distal end of the link rod230, in some embodiments, may be coupled to one end pivot 236 of rockingbeam 216, which is positioned vertically and perpendicularly under theproximate end of link rod 230. A stationary linear bearing 220 may bepositioned along coupling assembly 210 to further ensure linearlongitudinal motion of the piston rod 228 and thus ensuring linearlongitudinal motion of the piston 202. In an exemplary embodiment, linkrod 230 does not pass through linear bearing 220 to ensure that pistonrod 228 retains a substantially linear and longitudinal motion.

The coupling assemblies 210 and 212 change the alternating longitudinalmotion of respective pistons 202 and 204 to oscillatory motion of therocking beam 216. The delivered oscillatory motion is changed to therotational motion of the crankshaft 214 by the connecting rod 222,wherein one end of the connecting rod 222 is rotatably coupled to aconnecting pivot 240 positioned between an end pivot 232 and a rockerpivot 218 in the rocking beam 216, and another end of the connecting rod222 is rotatably coupled to crankpin 246. The rocker pivot 218 may bepositioned substantially at the midpoint between the end pivots 232 and236 and oscillatorily support the rocking beam 216 as a fulcrum, thusguiding the respective piston rods 224 and 228 to make sufficient linearmotion. In the exemplary embodiment, the crankshaft 214 is located abovethe rocking beam 216, but in other embodiments, the crankshaft 214 maybe positioned below the rocking beam 216 (as shown in FIGS. 5B and 5D)or in some embodiments, the crankshaft 214 is positioned to the side ofthe rocking beam 216, such that it still has a parallel axis to therocking beam 216.

Still referring to FIGS. 2-4, the rocking beam oscillates about therocker pivot 218, the end pivots 232 and 236 follow an arc path. Sincethe distal ends of the link rods 226 and 230 are connected to therocking beam 216 at pivots 232 and 236, the distal ends of the link rods226 and 230 also follow this arc path, resulting in an angular deviation242 and 244 from the longitudinal axis of motion of their respectivepistons 202 and 204. The coupling means 234 and 238 are configured suchthat any angular deviation 244 and 242 from the link rods 226 and 230experienced by the piston rods 224 and 228 is minimized. Essentially,the angular deviation 244 and 242 is absorbed by the coupling means 234and 238 so that the piston rods 224 and 228 maintain substantiallylinear longitudinal motion to reduce side loads on the pistons 204 and202. A stationary linear bearing 220 may also be placed inside thecylinder 208 or 206, or along coupling assemblies 212 or 210, to furtherabsorb any angular deviation 244 or 242 thus keeping the piston push rod224 or 228 and the piston 204 or 202 in linear motion along thelongitudinal axis of the piston 204 or 202.

Therefore, in view of reciprocating motion of pistons 202 and 204, it isnecessary to keep the motion of pistons 202 and 204 as close to linearas possible because the deviation 242 and 244 from longitudinal axis ofreciprocating motion of pistons 202 and 204 causes noise, reduction ofefficiency, increase of friction to the wall of cylinder, increase ofside-load, and low durability of the parts. The alignment of thecylinders 206 and 208 and the arrangement of crankshaft 214, piston rods224 and 228, link rods 226 and 230, and connecting rod 222, hence, mayinfluence on, amongst other things, the efficiency and/or the volume ofthe device. For the purpose of increasing the linearity of the pistonmotion as mentioned, the pistons (shown as 202 and 204 in FIGS. 2-4) arepreferably as close to the side of the respective cylinders 206 and 208as possible.

In another embodiment reducing angular deviation of link rods, link rods226 and 230 substantially linearly reciprocate along longitudinal axisof motion of respective pistons 204 and 202 to decrease the angulardeviation and thus to decrease the side load applied to each piston 204and 202. The angular deviation defines the deviation of the link rod 226or 230 from the longitudinal axis of the piston 204 or 202. Numerals 244and 242 designate the angular deviation of the link rods 226 and 230, asshown in FIG. 2. Therefore, the position of coupling assembly 212influences the angular displacement of the link rod 226, based on thelength of the distance between the end pivot 232 and the rocker pivot218 of the rocking beam 216. Thus, the position of the couplingassemblies may be such that the angular displacement of the link rod 226is reduced. For the link rod 230, the length of the coupling assembly210 also may be determined and placed to reduce the angular displacementof the link rod 230, based on the length of the distance between the endpivot 236 and the rocker pivot 218 of the rocking beam 216. Therefore,the length of the link rods 226 and 230, the length of couplingassemblies 212 and 210, and the length of the rocking beam 216 aresignificant parameters that greatly influence and/or determine theangular deviation of the link rods 226 and 230 as shown in FIG. 2.

The exemplary embodiment has a straight rocking beam 216 having the endpoints 232 and 236, the rocker pivot 218, and the connecting pivot 240along the same axis. However, in other embodiments, the rocking beam 216may be bent, such that pistons may be placed at angles to each other, asshown in FIGS. 5C and 5D.

Referring now to FIGS. 2-4 and FIGS. 7A-7B, in some embodiments of thecoupling assembly, the coupling assemblies 212 and 210, may include aflexible link rod that is axially stiff but flexible in the rocking beam216 plane of motion between link rods 226 and 230, and pistons 204 and202, respectively. In this embodiment, at least one portion, the flexure(shown as 700 in FIGS. 7A and 7B), of link rods 226 and 230 is elastic.The flexture 700 acts as a coupling means between the piston rod and thelink rod. The flexure 700 may absorb the crank-induced side loads of thepistons more effectively, thus allowing its respective piston tomaintain linear longitudinal movement inside the piston's cylinder. Thisflexure 700 allows small rotations in the plane of the rocking beam 216between the link rods 226 and 230 and pistons 204 or 202, respectively.Although depicted in this embodiment as flat, which increases theelasticity of the link rods 226 and 230, the flexure 700, in someembodiments, is not flat. The flexure 700 also may be constructed nearto the lower portion of the pistons or near to the distal end of thelink rods 226 and 230. The flexure 700, in one embodiment, may be madeof #D2 Tool Steel Hardened to 58-62 RC. In some embodiments, there maybe more than one flexure (not shown) on the link rod 226 or 230 toincrease the elasticity of the link rods.

In alternate embodiment, the axes of the pistons in each cylinderhousing may extend in different directions, as depicted in FIGS. 5C and5D. In the exemplary embodiment, the axes of the pistons in eachcylinder housing are substantially parallel and preferably substantiallyvertical, as depicted in FIGS. 2-4, and FIGS. 5A and 5B. FIGS. 5A-5Dinclude various embodiments of the rocking beam drive mechanismincluding like numbers as those shown and described with respect toFIGS. 2-4. It will be understood by those skilled in that art thatchanging the relative position of the connecting pivot 240 along therocking beam 216 will change the stroke of the pistons.

Accordingly, a change in the parameters of the relative position of theconnecting pivot 240 in the rocking beam 216 and the length of thepiston rods 224 and 228, link rods 230 and 226, rocking beam 216, andthe position of rocker pivot 218 will change the angular deviation ofthe link rods 226 and 230, the phasing of the pistons 204 and 202, andthe size of the device 300 in a variety of manner. Therefore, in variousembodiments, a wide range of piston phase angles and variable sizes ofthe engine may be chosen based on the modification of one or more ofthese parameters. In practice, the link rods 224 and 228 of theexemplary embodiment have substantially lateral movement within from−0.5 degree to +0.5 degree from the longitudinal axis of the pistons 204and 202. In various other embodiments, depending on the length of thelink rod, the angle may vary anywhere from approaching 0 degrees to 0.75degrees. However, in other embodiments, the angle may be higherincluding anywhere from approaching 0 to the approximately 20 degrees.As the link rod length increases, however, the crankcase/overall engineheight increases as well as the weight of the engine.

One feature of the exemplary embodiment is that each piston has its linkrod extending substantially to the attached piston rod so that it isformed as a coupling assembly. In one embodiment, the coupling assembly212 for the piston 204 includes a piston rod 224, a link rod 226, and acoupling means 234 as shown in FIG. 2. More specifically, one proximalend of piston rod 224 is attached to the lower portion of piston 204 andthe distal end piston rod 224 is connected to the proximate end of thelink rod 226 by the coupling means 234. The distal end of the link rod226 extends vertically to the end pivot 232 of the rocking beam 216. Asdescribed above, the coupling means 234 may be, but is not limited to, ajoint, hinge, coupling, or flexure or other means known in the art. Inthis embodiment, the ratio of the piston rod 224 and the link rod 226may determine the angular deviation of the link rod 226 as mentionedabove.

In one embodiment of the machine, an engine, such as a Stirling engine,employs more than one rocking beam drive on a crankshaft. Referring nowto FIG. 8, an unwrapped “four cylinder” rocking beam drive mechanism 800is shown. In this embodiment, the rocking beam drive mechanism has fourpistons 802, 804, 806, and 808 coupled to two rocking beam drives 810and 812. In the exemplary embodiment, rocking beam drive mechanism 800is used in a Stirling engine comprising at least four pistons 802, 804,806, and 808, positioned in a quadrilateral arrangement coupled to apair of rocking beam drives 810 and 812, wherein each rocking beam driveis connected to crankshaft 814. However, in other embodiments, theStirling cycle engine includes anywhere from 1-4 pistons, and in stillother embodiments, the Stirling cycle engine includes more than 4pistons. In some embodiments, rocking beam drives 810 and 812 aresubstantially similar to the rocking beam drives described above withrespect to FIGS. 2-4 (shown as 210 and 212 in FIGS. 2-4). Although inthis embodiment, the pistons are shown outside the cylinders, inpractice, the pistons would be inside cylinders.

Still referring to FIG. 8, in some embodiments, the rocking beam drivemechanism 800 has a single crankshaft 814 having a pair oflongitudinally spaced, radially and oppositely directed crank pins 816and 818 adapted for being journalled in a housing, and a pair of rockingbeam drives 810 and 812. Each rocking beam 820 and 822 is pivotallyconnected to rocker pivots 824 and 826, respectively, and to crankpins816 and 818, respectively. In the exemplary embodiment, rocking beams820 and 822 are coupled to a rocking beam shaft 828.

In some embodiments, a motor/generator may be connected to thecrankshaft in a working relationship. The motor may be located, in oneembodiment, between the rocking beam drives. In another embodiment, themotor may be positioned outboard. The term “motor/generator” is used tomean either a motor or a generator.

FIG. 9 shows one embodiment of crankshaft 814. Positioned on thecrankshaft is a motor/generator 900, such as a Permanent Magnetic (“PM”)generator. Motor/generator 900 may be positioned between, or inboard ofthe rocking beam drives (not shown, shown in FIG. 8 as 810 and 812), ormay be positioned outside, or outboard of, rocking beam drives 810 and812 at an end of crankshaft 814, as depicted by numeral 1000 in FIG.10A.

When motor/generator 900 is positioned between the rocking beam drives(not shown, shown in FIG. 8 as 810 and 812), the length ofmotor/generator 900 is limited to the distance between the rocking beamdrives. The diameter squared of motor/generator 900 is limited by thedistance between the crankshaft 814 and the rocking beam shaft 828.Because the capacity of motor/generator 900 is proportional to itsdiameter squared and length, these dimension limitations result in alimited-capacity “pancake” motor/generator 900 having relatively shortlength, and a relatively large diameter squared. The use of a “pancake”motor/generator 900 may reduce the overall dimension of the engine,however, the dimension limitations imposed by the inboard configurationresult in a motor/generator having limited capacity.

Placing motor/generator 900 between the rocking beam drives exposesmotor/generator 900 to heat generated by the mechanical friction of therocking beam drives. The inboard location of motor/generator 900 makesit more difficult to cool motor/generator 900, thereby increasing theeffects of heat produced by motor/generator 900 as well as heat absorbedby motor/generator 900 from the rocking beam drives. This may lead tooverheating, and ultimately failure of motor/generator 900.

Referring to both FIGS. 8 and 9, the inboard positioning ofmotor/generator 900 may also lead to an unequilateral configuration ofpistons 802, 804, 806, and 808, since pistons 802, 804, 806, and 808 arecoupled to rocking beam drives 810 and 812, respectively, and anyincrease in distance would also result in an increase in distancebetween pistons 802, 804, and pistons 806 and 808. An unequilateralarrangement of pistons may lead to inefficiencies in burner and heaterhead thermodynamic operation, which, in turn, may lead to a decrease inoverall engine efficiency. Additionally, an unequilateral arrangement ofpistons may lead to larger heater head and combustion chamberdimensions.

The exemplary embodiment of the motor/generator arrangement is shown inFIG. 10A. As shown in FIG. 10A, the motor/generator 1000 is positionedoutboard from rocking beam drives 1010 and 1012 (shown as 810 and 812 inFIG. 8) and at an end of crankshaft 1006. The outboard position allowsfor a motor/generator 1000 with a larger length and diameter squaredthan the “pancake” motor/generator described above (shown as 900 in FIG.9). As previously stated, the capacity of motor/generator 1000 isproportional to its length and diameter squared, and since outboardmotor/generator 1000 may have a larger length and diameter squared, theoutboard motor/generator 1000 configuration shown in FIG. 10A may allowfor the use of a higher capacity motor/generator in conjunction withengine.

By placing motor/generator 1000 outboard of drives 1010 and 1012 asshown in the embodiment in FIG. 10A, motor/generator 1000 is not exposedto heat generated by the mechanical friction of drives 1010 and 1012.Also, the outboard position of motor/generator 1000 makes it easier tocool the motor/generator, thereby allowing for more mechanical enginecycles per a given amount of time, which in turn allows for higheroverall engine performance.

Also, as motor/generator 1000 is positioned outside and not positionedbetween drives 1010 and 1012, rocking beam drives 1010 and 1012 may beplaced closer together thereby allowing the pistons which are coupled todrives 1010 and 1012 to be placed in an equilateral arrangement. In someembodiments, depending on the burner type used, particularly in the caseof a single burner embodiment, equilateral arrangement of pistons allowsfor higher efficiencies in burner and heater head thermodynamicoperation, which in turn allows higher overall engine performance.Equilateral arrangement of pistons also advantageously allows forsmaller heater head and combustion chamber dimensions.

Referring again to FIGS. 8 and 9, crankshaft 814 may have concentricends 902 and 904, which in one embodiment are crank journals, and invarious other embodiments, may be, but are not limited to, bearings.Each concentric end 902, 904 has a crankpin 816, 818 respectively, thatmay be offset from a crankshaft center axis. At least one counterweight906 may be placed at either end of crankshaft 814 (shown as 1006 in FIG.10A), to counterbalance any instability the crankshaft 814 mayexperience. This crankshaft configuration in combination with therocking beam drive described above allows the pistons (shown as 802,804, 806, and 808 in FIG. 8) to do work with one rotation of thecrankshaft 814. This characteristic will be further explained below. Inother embodiments, a flywheel (not shown) may be placed on crankshaft814 (shown as 1006 in FIG. 10A) to decrease fluctuations of angularvelocity for a more constant speed.

Still referring to FIGS. 8 and 9, in some embodiments, a cooler (notshown) may be also be positioned along the crankshaft 814 (shown as 1006in FIG. 10A) and rocking beam drives 810 and 812 (shown as 1010 and 1012in FIG. 10A) to cool the crankshaft 814 and rocking beam drives 810 and812. In some embodiments, the cooler may be used to cool the working gasin a cold chamber of a cylinder and may also be configured to cool therocking beam drive. Various embodiments of the cooler are discussed indetail below.

FIGS. 10A-10G depict some embodiments of various parts of the machine.As shown in this embodiment, crankshaft 1006 is coupled tomotor/generator 1000 via a motor/generator coupling assembly. Sincemotor/generator 1000 is mounted to crankcase 1008, pressurization ofcrankcase with a charge fluid may result in crankcase deformation, whichin turn may lead to misalignments between motor/generator 1000 andcrankshaft 1006 and cause crankshaft 1006 to deflect. Because rockingbeam drives 1010 and 1012 are coupled to crankshaft 1006, deflection ofcrankshaft 1006 may lead to failure of rocking beam drives 1010 and1012. Thus, in one embodiment of the machine, a motor/generator couplingassembly is used to couple the motor/generator 1000 to crankshaft 1006.The motor/generator coupling assembly accommodates differences inalignment between motor/generator 1000 and crankshaft 1006 which maycontribute to failure of rocking beam drives 1010 and 1012 duringoperation.

Still referring to FIGS. 10A-10G, in one embodiment, the motor/generatorcoupling assembly is a spline assembly that includes spline shaft 1004,sleeve rotor 1002 of motor/generator 1000, and crankshaft 1006. Splineshaft 1004 couples one end of crankshaft 1006 to sleeve rotor 1002.Sleeve rotor 1002 is attached to motor/generator 1000 by mechanicalmeans, such as press fitting, welding, threading, or the like. In oneembodiment, spline shaft 1004 includes a plurality of splines on bothends of the shaft. In other embodiments, spline shaft 1004 includes amiddle splineless portion 1014, which has a diameter smaller than theouter diameter or inner diameter of splined portions 1016 and 1018. Instill other embodiments, one end portion of the spline shaft 1016 hassplines that extend for a longer distance along the shaft than a secondend portion 1018 that also includes splines thereon.

In some embodiments, sleeve rotor 1002 includes an opening 1020 thatextends along a longitudinal axis of sleeve rotor 1002. The opening 1020is capable of receiving spline shaft 1004. In some embodiments, opening1020 includes a plurality of inner splines 1022 capable of engaging thesplines on one end of spline shaft 1004. The outer diameter 1028 ofinner splines 1022 may be larger than the outer diameter 1030 of thesplines on spline shaft 1004, such that the fit between inner splines1022 and the splines on spline shaft 1004 is loose (as shown in FIG.10E). A loose fit between inner splines 1022 and the splines on splineshaft 1004 contributes to maintain spline engagement between splineshaft 1004 and rotor sleeve 1002 during deflection of spline shaft 1004,which may be caused by crankcase pressurization. In other embodiments,longer splined portion 1016 of spline shaft 1004 may engage innersplines 1022 of rotor 1002.

Still referring to FIGS. 10A-10G, in some embodiments, crankshaft 1006has an opening 1024 on an end thereof, which is capable of receiving oneend of spline shaft 1004. Opening 1024 preferably includes a pluralityof inner splines 1026 that engage the splines on spline shaft 1004. Theouter diameter 1032 of inner splines 1026 may be larger than the outerdiameter 1034 of the splines on spline shaft 1004, such that the fitbetween inner splines 1026 and the splines on spline shaft 1004 is loose(as shown in FIG. 10F). As previously discussed, a loose fit betweeninner splines 1026 and the splines on spline shaft 1004 contributes tomaintain spline engagement between spline shaft 1004 and crankshaft 1006during deflection of spline shaft 1004, which may be caused by crankcasepressurization. The loose fit between the inner splines 1026 and 1022 onthe crankshaft 1006 and the sleeve rotor 1002 and the splines on thespline shaft 1004 may contribute to maintain deflection of spline shaft1004. This may allow misalignments between crankshaft 1006 and sleeverotor 1002. In some embodiments, shorter splined portion 1018 of splineshaft 1004 may engage opening 1024 of crankshaft 1006 thus preventingthese potential misalignments.

In some embodiments, opening 1020 of sleeve rotor 1002 includes aplurality of inner splines that extend the length of opening 1020. Thisarrangement contributes to spline shaft 1004 being properly insertedinto opening 1020 during assembly. This contributes to proper alignmentbetween the splines on spline shaft 1004 and the inner splines on sleeverotor 1002 being maintained.

Referring now to FIG. 4, one embodiment of the engine is shown. Here thepistons 202 and 204 of engine 300 operate between a hot chamber 404 anda cold chamber 406 of cylinders 206 and 208 respectively. Between thetwo chambers there may be a regenerator 408. The regenerator 408 mayhave variable density, variable area, and, in some embodiments, is madeof wire. The varying density and area of the regenerator may be adjustedsuch that the working gas has substantially uniform flow across theregenerator 408. Various embodiments of the regenerator 408 arediscussed in detail below, and in U.S. Pat. No. 6,591,609, issued Jul.17, 2003, to Kamen et al., and No. 6,862,883, issued Mar. 8, 2005, toKamen et al., which are herein incorporated by reference in theirentireties. When the working gas passes through the hot chamber 404, aheater head 410 may heat the gas causing the gas to expand and pushpistons 202 and 204 towards the cold chamber 406, where the gascompresses. As the gas compresses in the cold chamber 406, pistons 202and 204 may be guided back to the hot chamber to undergo the Stirlingcycle again. The heater head 410 may be a pin head (as shown in FIGS.52A through 53B), a fin head (as shown in FIGS. 56A through 56C), afolded fin head (as shown in FIGS. 56A through 56C), heater tubes asshown in FIG. 4 (also shown as 2904 in FIG. 29), or any other heaterhead embodiment known, including, but not limited to, those describedbelow. Various embodiments of heater head 410 are discussed in detailbelow, and in U.S. Pat. No. 6,381,958, issued May 7, 2002, to Kamen etal., U.S. Pat. No. 6,543,215, issued Apr. 8, 2003, to Langenfeld et al.,U.S. Pat. No. 6,966,182, issued Nov. 22, 2005, to Kamen et al, and U.S.Pat. No. 7,308,787, issued Dec. 18, 2007, to LaRocque et al., which areherein incorporated by reference in their entireties.

In some embodiments, a cooler 412 may be positioned alongside cylinders206 and 208 to further cool the gas passing through to the cold chamber406. Various embodiments of cooler 412 are discussed in detail in theproceeding sections, and in U.S. Pat. No. 7,325,399, issued Feb. 5,2008, to Strimling et al, which is herein incorporated by reference inits entirety.

In some embodiments, at least one piston seal 414 may be positioned onpistons 202 and 204 to seal the hot section 404 off from the coldsection 406. Additionally, at least one piston guide ring 416 may bepositioned on pistons 202 and 204 to help guide the pistons' motion intheir respective cylinders. Various embodiments of piston seal 414 andguide ring 416 are described in detail below, and in U.S. patentapplication Ser. No. 10/175,502, filed Jun. 19, 2002, published Feb. 6,2003 (now abandoned), which is herein incorporated by reference in itsentirety.

In some embodiments, at least one piston rod seal 418 may be placedagainst piston rods 224 and 228 to prevent working gas from escapinginto the crankcase 400, or alternatively into airlock space 420. Thepiston rod seal 418 may be an elastomer seal, or a spring-loaded seal.Various embodiments of the piston rod seal 418 are discussed in detailbelow.

In some embodiments, the airlock space may be eliminated, for example,in the rolling diaphragm and/or bellows embodiments described in moredetail below. In those cases, the piston rod seals 224 and 228 seal theworking space from the crankcase.

In some embodiments, at least one rolling diaphragm/bellows 422 may belocated along piston rods 224 and 228 to prevent airlock gas fromescaping into the crankcase 400. Various embodiments of rollingdiaphragm 422 are discussed in more detail below.

Although FIG. 4 shows a cross section of engine 300 depicting only twopistons and one rocking beam drive, it is to be understood that theprinciples of operation described herein may apply to a four cylinder,double rocking beam drive engine, as designated generally by numeral 800in FIG. 8.

Piston Operation

Referring now to FIGS. 8 and 11, FIG. 11 shows the operation of pistons802, 804, 806, and 808 during one revolution of crankshaft 814. With a ¼revolution of crankshaft 814, piston 802 is at the top of its cylinder,otherwise known as top dead center, piston 806 is in upward midstroke,piston 804 is at the bottom of its cylinder, otherwise known as bottomdead center, and piston 808 is in downward midstroke. With a ½revolution of crankshaft 814, piston 802 is in downward midstroke,piston 806 is at top dead center, piston 804 is in upward midstroke, andpiston 808 is at bottom dead center. With ¾ revolution of crankshaft814, piston 802 is at bottom dead center, piston 806 is in downwardmidstroke, piston 804 is at top dead center, and piston 808 is in upwardmidstroke. Finally, with a full revolution of crankshaft 814, piston 802is in upward midstroke, piston 806 is at bottom dead center, piston 804is in downward midstroke, and piston 808 is at top dead center. Duringeach ¼ revolution, there is a 90 degree phase difference between pistons802 and 806, a 180 degree phase difference between pistons 802 and 804,and a 270 degree phase difference between pistons 802 and 808. FIG. 12Aillustrates the relationship of the pistons being approximately 90degrees out of phase with the preceding and succeeding piston.Additionally, FIG. 11 shows the exemplary embodiment machine means oftransferring work. Thus, work is transferred from piston 802 to piston806 to piston 804 to piston 808 so that with a full revolution ofcrankshaft 814, all pistons have exerted work by moving from the top tothe bottom of their respective cylinders.

Referring now to FIG. 11, together with FIGS. 12A-12C, illustrate the 90degree phase difference between the pistons in the exemplary embodiment.Referring now to FIG. 12A, although the cylinders are shown in a linearpath, this is for illustration purposes only. In the exemplaryembodiment of a four cylinder Stirling cycle machine, the flow path ofthe working gas contained within the cylinder working space follows afigure eight pattern. Thus, the working spaces of cylinders 1200, 1202,1204, and 1206 are connected in a figure eight pattern, for example,from cylinder 1200 to cylinder 1202 to cylinder 1204 to cylinder 1208,the fluid flow pattern follows a figure eight. Still referring to FIG.12A, an unwrapped view of cylinders 1200, 1202, 1204, and 1206, takenalong the line B-B (shown in FIG. 12C) is illustrated. The 90 degreephase difference between pistons as described above allows for theworking gas in the warm section 1212 of cylinder 1204 to be delivered tothe cold section 1222 of cylinder 1206. As piston 802 and 808 are 90degrees out of phase, the working gas in the warm section 1214 ofcylinder 1206 is delivered to the cold section 1216 of cylinder 1200. Aspiston 802 and piston 806 are also 90 degrees out of phase, the workinggas in the warm section 1208 of cylinder 1200 is delivered to the coldsection 1218 of cylinder 1202. And as piston 804 and piston 806 are also90 degrees out of phase, so the working gas in the warm section 1210 ofcylinder 1202 is delivered to the cold section 1220 of cylinder 1204.Once the working gas of a warm section of a first cylinder enters thecold section of a second cylinder, the working gas begins to compress,and the piston within the second cylinder, in its down stroke,thereafter forces the compressed working gas back through a regenerator1224 and heater head 1226 (shown in FIG. 12B), and back into the warmsection of the first cylinder. Once inside the warm section of the firstcylinder, the gas expands and drives the piston within that cylinderdownward, thus causing the working gas within the cold section of thatfirst cylinder to be driven through the preceding regenerator and heaterhead, and into the cylinder. This cyclic transmigration characteristicof working gas between cylinders 1200, 1202, 1204, and 1206 is possiblebecause pistons 802, 804, 806, and 808 are connected, via drives 810 and812, to a common crankshaft 814 (shown in FIG. 11), in such a way thatthe cyclical movement of each piston is approximately 90 degrees inadvance of the movement of the proceeding piston, as depicted in FIG.12A.

Rolling Diaphragm, Metal Bellows, Airlock, and Pressure Regulator

In some embodiments of the Stirling cycle machine, lubricating fluid isused. To prevent the lubricating fluid from escaping the crankcase, aseal is used.

Referring now to FIGS. 13A-15, some embodiments of the Stirling cyclemachine include a fluid lubricated rocking beam drive that utilizes arolling diaphragm 1300 positioned along the piston rod 1302 to preventlubricating fluid from escaping the crankcase, not shown, but thecomponents that are housed in the crankcase are represented as 1304, andentering areas of the engine that may be damaged by the lubricatingfluid. It is beneficial to contain the lubricating fluid for iflubricating fluid enters the working space, not shown, but thecomponents that are housed in the working space are represented as 1306,it would contaminate the working fluid, come into contact with theregenerator 1308, and may clog the regenerator 1308. The rollingdiaphragm 1300 may be made of an elastomer material, such as rubber orrubber reinforced with woven fabric or non-woven fabric to providerigidity. The rolling diaphragm 1300 may alternatively be made of othermaterials, such as fluorosilicone or nitrile with woven fabric ornon-woven fabric. The rolling diaphragm 1300 may also be made of carbonnanotubes or chopped fabric, which is non-woven fabric with fibers ofpolyester or KEVLAR®, for example, dispersed in an elastomer. In thesome embodiments, the rolling diaphragm 1300 is supported by the topseal piston 1328 and the bottom seal piston 1310. In other embodiments,the rolling diaphragm 1300 as shown in FIG. 13A is supported via notchesin the top seal piston 1328.

In some embodiments, a pressure differential is placed across therolling diaphragm 1300 such that the pressure above the seal 1300 isdifferent from the pressure in the crankcase 1304. This pressuredifferential inflates seal 1300 and allows seal 1300 to act as a dynamicseal as the pressure differential ensures that rolling diaphragmmaintains its form throughout operation. FIG. 13A, and FIGS. 13C-13Hillustrate how the pressure differential effects the rolling diaphragm.The pressure differential causes the rolling diaphragm 1300 to conformto the shape of the bottom seal piston 1310 as it moves with the pistonrod 1302, and prevents separation of the seal 1300 from a surface of thepiston 1310 during operation. Such separation may cause seal failure.The pressure differential causes the rolling diaphragm 1300 to maintainconstant contact with the bottom seal piston 1310 as it moves with thepiston rod 1302. This occurs because one side of the seal 1300 willalways have pressure exerted on it thereby inflating the seal 1300 toconform to the surface of the bottom seal piston 1310. In someembodiments, the top seal piston 1328 ‘rolls over’ the corners of therolling diaphragm 1300 that are in contact with the bottom seal piston1310, so as to further maintain the seal 1300 in contact with the bottomseal piston 1310. In the exemplary embodiment, the pressure differentialis in the range of 10 to 15 PSI. The smaller pressure in the pressuredifferential is preferably in crankcase 1304, so that the rollingdiaphragm 1300 may be inflated into the crankcase 1304. However, inother embodiments, the pressure differential may have a greater orsmaller range of value.

The pressure differential may be created by various methods including,but not limited to, the use of the following: a pressurized lubricationsystem, a pneumatic pump, sensors, an electric pump, by oscillating therocking beam to create a pressure rise in the crankcase 1304, bycreating an electrostatic charge on the rolling diaphragm 1300, or othersimilar methods. In some embodiments, the pressure differential iscreated by pressurizing the crankcase 1304 to a pressure that is belowthe mean pressure of the working space 1306. In some embodiments thecrankcase 1304 is pressurized to a pressure in the range of 10 to 15 PSIbelow the mean pressure of the working space 1306, however, in variousother embodiments, the pressure differential may be smaller or greater.Further detail regarding the rolling diaphragm is included below.

Referring now to FIGS. 13C, 13G, and 13H, however, another embodiment ofthe Stirling machine is shown, wherein airlock space 1312 is locatedbetween working space 1306 and crankcase 1304. Airlock space 1312maintains a constant volume and pressure necessary to create thepressure differential necessary for the function of rolling diaphragm1300 as described above. In one embodiment, airlock 1312 is notabsolutely sealed off from working space 1306, so the pressure ofairlock 1312 is equal to the mean pressure of working space 1306. Thus,in some embodiments, the lack of an effective seal between the workingspace and the crankcase contributes to the need for an airlock space.Thus, the airlock space, in some embodiments, may be eliminated by amore efficient and effective seal.

During operation, the working space 1306 mean pressure may vary so as tocause airlock 1312 mean pressure to vary as well. One reason thepressure may tend to vary is that during operation the working space mayget hotter, which in turn may increase the pressure in the workingspace, and consequently in the airlock as well since the airlock andworking space are in fluid communication. In such a case, the pressuredifferential between airlock 1312 and crankcase 1304 will also vary,thereby causing unnecessary stresses in rolling diaphragms 1300 that maylead to seal failure. Therefore, some embodiments of the machine, themean pressure within airlock 1312 is regulated so as to maintain aconstant desired pressure differential between airlock 1312 andcrankcase 1304, and ensuring that rolling diaphragms 1300 stay inflatedand maintains their form. In some embodiments, a pressure transducer isused to monitor and manage the pressure differential between the airlockand the crankcase, and regulate the pressure accordingly so as tomaintain a constant pressure differential between the airlock and thecrankcase. Various embodiments of the pressure regulator that may beused are described in further detail below, and in U.S. Pat. No.7,310,945, issued Dec. 25, 2007, to Gurski et al., which is hereinincorporated by reference in its entirety.

A constant pressure differential between the airlock 1312 and crankcase1304 may be achieved by adding or removing working fluid from airlock1312 via a pump or a release valve. Alternatively, a constant pressuredifferential between airlock 1312 and crankcase 1304 may be achieved byadding or removing working fluid from crankcase 1304 via a pump or arelease valve. The pump and release valve may be controlled by thepressure regulator. Working fluid may be added to airlock 1312 (orcrankcase 1304) from a separate source, such as a working fluidcontainer, or may be transferred over from crankcase 1304. Shouldworking fluid be transferred from crankcase 1304 to airlock 1312, it maybe desirable to filter the working fluid before passing it into airlock1312 so as to prevent any lubricant from passing from crankcase 1304into airlock 1312, and ultimately into working space 1306, as this mayresult in engine failure.

In some embodiments of the machine, crankcase 1304 may be charged with afluid having different thermal properties than the working fluid. Forexample, where the working gas is helium or hydrogen, the crankcase maybe charged with argon. Thus, the crankcase is pressurized. In someembodiments, helium is used, but in other embodiments, any inert gas, asdescribed herein, may be used. Thus, the crankcase is a wet pressurizedcrankcase in the exemplary embodiment. In other embodiments where alubricating fluid is not used, the crankcase is not wet.

In the exemplary embodiments, rolling diaphragms 1300 do not allow gasor liquid to pass through them, which allows working space 1306 toremain dry and crankcase 1304 to be wet sumped with a lubricating fluid.Allowing a wet sump crankcase 1304 increases the efficiency and life ofthe engine as there is less friction in rocking beam drives 1316. Insome embodiments, the use of roller bearings or ball bearings in drives1316 may also be eliminated with the use of lubricating fluid androlling diaphragms 1300. This may further reduce engine noise andincrease engine life and efficiency.

FIGS. 14A-14E show cross sections of various embodiments of the rollingdiaphragm (shown as 1400, 1410, 1412, 1422 and 1424) configured to bemounted between top seal piston and bottom seal piston (shown as 1328and 1310 in FIGS. 13A and 13H), and between a top mounting surface and abottom mounting surface (shown as 1320 and 1318 in FIG. 13A). In someembodiments, the top mounting surface may be the surface of an airlockor working space, and the bottom mounting surface may be the surface ofa crankcase.

FIG. 14A shows one embodiment of the rolling diaphragm 1400, where therolling diaphragm 1400 includes a flat inner end 1402 that may bepositioned between a top seal piston and a bottom seal piston, so as toform a seal between the top seal piston and the bottom seal piston. Therolling diaphragm 1400 also includes a flat outer end 1404 that may bepositioned between a top mounting surface and a bottom mounting surface,so as to form a seal between the top mounting surface and the bottommounting surface. FIG. 14B shows another embodiment of the rollingdiaphragm, wherein rolling diaphragm 1410 may include a plurality ofbends 1408 leading up to flat inner end 1406 to provide for additionalsupport and sealing contact between the top seal piston and the bottomseal piston. FIG. 14C shows another embodiment of the rolling diaphragm,wherein rolling diaphragm 1412 includes a plurality of bends 1416leading up to flat outer end 1414 to provide for additional support andsealing contact between the top mounting surface and the bottom mountingsurface.

FIG. 14D shows another embodiment of the rolling diaphragm where rollingdiaphragm 1422 includes a bead along an inner end 1420 thereof, so as toform an ‘o-ring’ type seal between a top seal piston and a bottom sealpiston, and a bead along an outer end 1418 thereof, so as to form an‘o-ring’ type seal between a bottom mounting surface and a top mountingsurface. FIG. 14E shows another embodiment of the rolling diaphragm,wherein rolling diaphragm 1424 includes a plurality of bends 1428leading up to beaded inner end 1426 to provide for additional supportand sealing contact between the top seal piston and the bottom sealpiston. Rolling diaphragm 1424 may also include a plurality of bends1430 leading up to beaded outer end 1432 to provide for additionalsupport and sealing contact between the top seal piston and the bottomseal piston.

Although FIGS. 14A through 14E depict various embodiments of the rollingdiaphragm, it is to be understood that rolling diaphragms may be held inplace by any other mechanical means known in the art.

Referring now to FIG. 15A, a cross section shows one embodiment of therolling diaphragm embodiment. A metal bellows 1500 is positioned along apiston rod 1502 to seal off a crankcase (shown as 1304 in FIG. 13G) froma working space or airlock (shown as 1306 and 1312 in FIG. 13G). Metalbellows 1500 may be attached to a top seal piston 1504 and a stationarymounting surface 1506. Alternatively, metal bellows 1500 may be attachedto a bottom seal piston (not shown), and a top stationary mountingsurface. In one embodiment the bottom stationary mounting surface may bea crankcase surface or an inner airlock or working space surface, andthe top stationary mounting surface may be an inner crankcase surface,or an outer airlock or working space surface. Metal bellows 1500 may beattached by welding, brazing, or any mechanical means known in the art.

FIGS. 15B-15G depict a perspective cross sectional view of variousembodiments of the metal bellows, wherein the metal bellows is a weldedmetal bellows 1508. In some embodiments of the metal bellows, the metalbellows is preferably a micro-welded metal bellows. In some embodiments,the welded metal bellows 1508 includes a plurality of diaphragms 1510,which are welded to each other at either an inner end 1512 or an outerend 1514, as shown in FIGS. 15C and 15D. In some embodiments, diaphragms1510 may be crescent shaped 1516, flat 1518, rippled 1520, or any othershape known in the art.

Additionally, the metal bellows may alternatively be formed mechanicallyby means such as die forming, hydroforming, explosive hydroforming,hydramolding, or any other means known in the art.

The metal bellows may be made of any type of metal, including but notlimited to, steel, stainless steel, stainless steel 374, AM-350stainless steel, Inconel, Hastelloy, Haynes, titanium, or any otherhigh-strength, corrosion-resistant material.

In one embodiment, the metal bellows used are those available fromSenior Aerospace Metal Bellows Division, Sharon, Mass., or American BOA,Inc., Cumming, Ga.

Rolling Diaphragm and/or Bellows Embodiments

Various embodiments of the rolling diaphragm and/or bellows, whichfunction to seal, are described above. Further embodiments will beapparent to those of skill in the art based on the description above andthe additional description below relating to the parameters of therolling diaphragm and/or bellows.

In some embodiments, the pressure atop the rolling diaphragm or bellows,in the airlock space or airlock area (both terms are usedinterchangeably), is the mean-working-gas pressure for the machine,which, in some embodiments is an engine, while the pressure below therolling diaphragm and/or bellows, in the crankcase area, isambient/atmospheric pressure. In these embodiments, the rollingdiaphragm and/or bellows is required to operate with as much as 3000 psiacross it (and in some embodiments, up to 1500 psi or higher). In thiscase, the rolling diaphragm and/or bellows seal forms the working gas(helium, hydrogen, or otherwise) containment barrier for the machine(engine in the exemplary embodiment). Also, in these embodiments, theneed for a heavy, pressure-rated, structural vessel to contain thebottom end of the engine is eliminated, since it is now required tosimply contain lubricating fluid (oil is used as a lubricating fluid inthe exemplary embodiment) and air at ambient pressure, like aconventional internal combustion (“IC”) engine.

The capability to use a rolling diaphragm and/or bellows seal with suchan extreme pressure across it depends on the interaction of severalparameters. Referring now to FIG. 15H, an illustration of the actualload on the rolling diaphragm or bellows material is shown. As shown,the load is a function of the pressure differential and the annular gaparea for the installed rolling diaphragm or bellows seal.

Region 1 represents the portions of the rolling diaphragm and/or bellowsthat are in contact with the walls formed by the piston and cylinder.The load is essentially a tensile load in the axial direction, due tothe pressure differential across the rolling diaphragm and/or bellows.This tensile load due to the pressure across the rolling diaphragmand/or bellows can be expressed as:L _(t) =P _(d) *A _(a)Where

L_(t)=Tensile Load and

P_(d)=Pressure Differential

A_(a)=Annular Area

andA _(a) =p/4*(D ² −d ²)Where

D=Cylinder Bore and

d=Piston Diameter

The tensile component of stress in the bellows material can beapproximated as:S _(t) =L _(t)/(p*(D+d)*t _(b))Which reduces to:S _(t) =P _(d)/4*(D−d)/tbLater, we will show the relationship of radius of convolution, R_(c), toCylinder bore (D) and Piston Diameter (d) to be defined as:R _(c)=(D−d)/4So, this formula for St reduces to its final form:S _(t) =P _(d) *R _(c) /t _(b)Where

t_(b)=thickness of bellows material

Still referring to FIG. 15H, Region 2 represents the convolution. As therolling diaphragm and/or bellows material turns the corner, in theconvolution, the hoop stress imposed on the rolling diaphragm and/orbellows material may be calculated. For the section of the bellowsforming the convolution, the hoop component of stress can be closelyapproximated as:S _(h) =P _(d) *R _(c) /t _(b)

The annular gap that the rolling diaphragm and/or bellows rolls withinis generally referred to as the convolution area. The rolling diaphragmand/or bellows fatigue life is generally limited by the combined stressfrom both the tensile (and hoop) load, due to pressure differential, aswell as the fatigue due to the bending as the fabric rolls through theconvolution. The radius that the fabric takes on during this ‘rolling’is defined here as the radius of convolution, Rc.R _(c)=(D−d)/4

The bending stress, Sb, in the rolling diaphragm and/or bellows materialas it rolls through the radius of convolution, Rc, is a function of thatradius, as well as the thickness of the materials in bending. For afiber-reinforced material, the stress in the fibers themselves (duringthe prescribed deflection in the exemplary embodiments) is reduced asthe fiber diameter decreases. The lower resultant stress for the samelevel of bending allows for an increased fatigue life limit. As thefiber diameter is further reduced, flexibility to decrease the radius ofconvolution Rc is achieved, while keeping the bending stress in thefiber under its endurance limit. At the same time, as Rc decreases, thetensile load on the fabric is reduced since there is less unsupportedarea in the annulus between the piston and cylinder. The smaller thefiber diameter, the smaller the minimum Rc, the smaller the annulararea, which results in a higher allowable pressure differential.

For bending around a prescribed radius, the bending moment isapproximated by:M=E*I/RWhere:

M=Bending Moment

E=Elastic Modulus

I=Moment of Inertia

R=Radius of Bend

Classical bending stress, S_(b), is calculated as:S _(b) =M*Y/IWhere:

Y=Distance above neutral axis of bending

Substituting yields:S _(b)=(E*I/R)*Y/IS _(b) =E*Y/RAssuming bending is about a central neutral axis:Y _(max) =t _(b)/2S _(b) =E*t _(b)/(2*R)

In some embodiments, rolling diaphragm and/or bellows designs for highcycle life are based on geometry where the bending stress imposed iskept about one order of magnitude less than the pressure-based loading(hoop and axial stresses). Based on the equation: Sb=E*tb/(2*R), it isclear that minimizing tb in direct proportion to Rc should not increasethe bending stress. The minimum thickness for the exemplary embodimentsof the rolling diaphragm and/or bellows material or membrane is directlyrelated to the minimum fiber diameter that is used in the reinforcementof the elastomer. The smaller the fibers used, the smaller resultant Rcfor a given stress level.

Another limiting component of load on the rolling diaphragm and/orbellows is the hoop stress in the convolution (which is theoreticallythe same in magnitude as the axial load while supported by the piston orcylinder). The governing equation for that load is as follows:Sh=Pd*Rc/tb

Thus, if Rc is decreased in direct proportion to tb, then there is noincrease of stress on the membrane in this region. However, if thisratio is reduced in a manner that decreases Rc to a greater ratio thantb then parameters must be balanced. Thus, decreasing tb with respect toRc requires the rolling diaphragm and/or bellows to carry a heavierstress due to pressure, but makes for a reduced stress level due tobending. The pressure-based load is essentially constant, so this may befavorable—since the bending load is cyclic, therefore it is the bendingload component that ultimately limits fatigue life.

For bending stress reduction, tb ideally should be at a minimum, and Rcideally should be at a maximum. E ideally is also at a minimum. For hoopstress reduction, Rc ideally is small, and tb ideally is large.

Thus, the critical parameters for the rolling diaphragm and/or bellowsmembrane material are:

E, Elastic Modulus of the membrane material;

tb, membrane thickness (and/or fiber diameter);

Sut, Ultimate tensile strength of the rolling diaphragm and/or bellows;and

Slcf, The limiting fatigue strength of the rolling diaphragm and/orbellows.

Thus, from E, tb and Sut, the minimum acceptable Rc may be calculated.Next, using Rc, Slcf, and tb, the maximum Pd may be calculates. Rc maybe adjusted to shift the bias of load (stress) components between thesteady state pressure stress and the cyclic bending stress. Thus, theideal rolling diaphragm and/or bellows material is extremely thin,extremely strong in tension, and very limber in flexion.

Thus, in some embodiments, the rolling diaphragm and/or bellows material(sometimes referred to as a “membrane”), is made from carbon fibernanotubes. However, additional small fiber materials may also be used,including, but not limited to nanotube fibers that have been braided,nanotube untwisted yarn fibers, or any other conventional materials,including but not limited to KEVLAR, glass, polyester, synthetic fibersand any other material or fiber having a desirable diameter and/or otherdesired parameters as described in detail above.

Piston Seals and Piston Rod Seals

Referring now to FIG. 13G, an embodiment of the machine is shown whereinan engine 1326, such as a Stirling cycle engine, includes at least onepiston rod seal 1314, a piston seal 1324, and a piston guide ring 1322,(shown as 1616 in FIG. 16). Various embodiments of the piston seal 1324and the piston guide ring 1322 are further discussed below, and in U.S.patent application Ser. No. 10/175,502 (now abandoned), which, asmentioned before, is incorporated by reference.

FIG. 16 shows a partial cross section of the piston 1600, driven alongthe central axis 1602 of cylinder, or the cylinder 1604. The piston seal(shown as 1324 in FIG. 13G) may include a seal ring 1606, which providesa seal against the contact surface 1608 of the cylinder 1604. Thecontact surface 1608 is typically a hardened metal (preferably 58-62 RC)with a surface finish of 12 RMS or smoother. The contact surface 1608may be metal which has been case hardened, such as 8260 hardened steel,which may be easily case hardened and may be ground and/or honed toachieve a desired finish. The piston seal may also include a backingring 1610, which is sprung to provide a thrust force against the sealring 1606 thereby providing sufficient contact pressure to ensuresealing around the entire outward surface of the seal ring 1606. Theseal ring 1606 and the backing ring 1610 may together be referred to asa piston seal composite ring. In some embodiments, the at least onepiston seal may seal off a warm portion of cylinder 1604 from a coldportion of cylinder 1604.

Referring now to FIG. 17, some embodiments include a piston rod seal(shown as 1314 in FIG. 13G) mounted in the piston rod cylinder wall1700, which, in some embodiments, may include a seal ring 1706, whichprovides a seal against the contact surface 1708 of the piston rod 1604(shown as 1302 in FIG. 13G). The contact surface 1708 in someembodiments is a hardened metal (preferably 58-62 RC) with a surfacefinish of 12 RMS or smoother. The contact surface 1708 may be metalwhich has been case hardened, such as 8260 hardened steel, which may beeasily case hardened and may be ground and/or honed to achieve a desiredfinish. The piston seal may also include a backing ring 1710, which issprung to provide a radial or hoop force against the seal ring 1706thereby providing sufficient contact hoop stress to ensure sealingaround the entire inward surface of seal ring 1706. The seal ring 1706and the backing ring 1710 may together be referred to as a piston rodseal composite ring.

In some embodiments, the seal ring and the backing ring may bepositioned on a piston rod, with the backing exerting an outwardpressure on the seal ring, and the seal ring may come into contact witha piston rod cylinder wall 1702. These embodiments require a largerpiston rod cylinder length than the previous embodiment. This is becausethe contact surface on the piston rod cylinder wall 1702 will be longerthan in the previous embodiment, where the contact surface 1708 lies onthe piston rod itself. In yet another embodiment, piston rod seals maybe any functional seal known in the art including, but not limited to,an o-ring, a graphite clearance seal, graphite piston in a glasscylinder, or any air pot, or a spring energized lip seal. In someembodiments, anything having a close clearance may be used, in otherembodiments, anything having interference, for example, a seal, is used.In the exemplary embodiment, a spring energized lip seal is used. Anyspring energized lip seal may be used, including those made by BAL SEALEngineering, Inc., Foothill Ranch, Calif. In some embodiments, the sealused is a BAL SEAL Part Number X558604.

The material of the seal rings 1606 and 1706 is chosen by considering abalance between the coefficient of friction of the seal rings 1606 and1706 against the contact surfaces 1608 and 1708, respectively, and thewear on the seal rings 1606 and 1706 it engenders. In applications inwhich piston lubrication is not possible, such as at the high operatingtemperatures of a Stirling cycle engine, the use of engineering plasticrings is used. The embodiments of the composition include a nylon matrixloaded with a lubricating and wear-resistant material. Examples of suchlubricating materials include PTFE/silicone, PTFE, graphite, etc.Examples of wear-resistant materials include glass fibers and carbonfibers. Examples of such engineering plastics are manufactured by LNPEngineering Plastics, Inc. of Exton, Pa. Backing rings 1610 and 1710 ispreferably metal.

The fit between the seal rings 1606 and 1706 and the seal ring grooves1612 and 1712, respectively, is preferably a clearance fit (about0.002″), while the fit of the backing rings 1610 and 1710 is preferablya looser fit, of the order of about 0.005″ in some embodiments. The sealrings 1606 and 1706 provide a pressure seal against the contact surfaces1608 and 1708, respectively, and also one of the surfaces 1614 and 1714of the seal ring grooves 1612 and 1712, respectively, depending on thedirection of the pressure difference across the rings 1606 and 1706 andthe direction of the piston 1600 or the piston rod 1704 travel.

FIGS. 18A and 18B show that if the backing ring 1820 is essentiallycircularly symmetrical, but for the gap 1800, it will assume, uponcompression, an oval shape, as shown by the dashed backing ring 1802.The result may be an uneven radial or hoop force (depicted by arrows1804) exerted on the seal ring (not shown, shown as 1606 and 1706 inFIGS. 16 and 17), and thus an uneven pressure of the seal rings againstthe contact surfaces (not shown, shown as 1608 and 1708 in FIGS. 16 and17) respectively, causing uneven wear of the seal rings and in somecases, failure of the seals.

A solution to the problem of uneven radial or hoop force exerted by thepiston seal backing ring 1820, in accordance with an embodiment, is abacking ring 1822 having a cross-section varying with circumferentialdisplacement from the gap 1800, as shown in FIGS. 18C and 18D. Atapering of the width of the backing ring 1822 is shown from theposition denoted by numeral 1806 to the position denoted by numeral1808. Also shown in FIGS. 18C and 18D is a lap joint 1810 providing forcircumferential closure of the seal ring 1606. As some seals will wearsignificantly over their lifetime, the backing ring 1822 should providean even pressure (depicted by numeral 1904 in FIG. 19B) of a range ofmovement. The tapered backing ring 1822 shown in FIGS. 18C and 18D mayprovide this advantage.

FIGS. 19A and 19B illustrate another solution to the problem of unevenradial or hoop force of the piston seal ring against the pistoncylinder, in accordance with some embodiments. As shown in FIG. 19A,backing ring 1910 is fashioned in an oval shape, so that uponcompression within the cylinder, the ring assumes the circular shapeshown by dashed backing ring 1902. A constant contact pressure betweenthe seal ring and the cylinder contact surface may thus be provided byan even radial force 1904 of backing ring 1902, as shown in FIG. 19B.

A solution to the problem of uneven radial or hoop force exerted by thepiston rod seal backing ring, in accordance with some embodiments, is abacking ring 1824 having a cross-section varying with circumferentialdisplacement from gap 1812, as shown in FIGS. 18E and 18F. A tapering ofthe width of backing ring 1824 is shown from the position denoted bynumeral 1814 to the position denoted by numeral 1816. Also shown inFIGS. 18E and 18F is a lap joint 1818 providing for circumferentialclosure of seal ring 1706. As some seals will wear significantly overtheir lifetime, backing ring 1824 should provide an even pressure(depicted by numeral 2004 in FIG. 20B) of a range of movement. Thetapered backing ring 1824 shown in FIGS. 18E and 18F may provide thisadvantage.

FIGS. 20A and 20B illustrate another solution to the problem of unevenradial or hoop force of the piston rod seal ring against the piston rodcontact surface, in accordance with some embodiments. As shown in FIG.20A, backing ring (shown by dashed backing ring 2000) is fashioned as anoval shape, so that upon expansion within the cylinder, the ring assumesthe circular shape shown by backing ring 2002. A constant contactpressure between the seal ring 1706 and the cylinder contact surface maythus be provided by an even radial thrust force 2004 of backing ring2002, as shown in FIG. 20B.

Referring again to FIG. 16, at least one guide ring 1616 may also beprovided, in accordance with some embodiments, for bearing any side loadon piston 1600 as it moves up and down the cylinder 1604. Guide ring1616 is also preferably fabricated from an engineering plastic materialloaded with a lubricating material. A perspective view of guide ring1616 is shown in FIG. 21. An overlapping joint 2100 is shown and may bediagonal to the central axis of guide ring 1616.

Lubricating Fluid Pump and Lubricating Fluid Passageways

Referring now to FIG. 22, a representative illustration of oneembodiment of the engine 2200 for the machine is shown having a rockingbeam drive 2202 and lubricating fluid 2204. In some embodiments, thelubricating fluid is oil. The lubricating fluid is used to lubricateengine parts in the crankcase 2206, such as hydrodynamic pressure fedlubricated bearings. Lubricating the moving parts of the engine 2200serves to further reduce friction between engine parts and furtherincrease engine efficiency and engine life. In some embodiments,lubricating fluid may be placed at the bottom of the engine, also knownas an oil sump, and distributed throughout the crankcase. Thelubricating fluid may be distributed to the different parts of theengine 2200 by way of a lubricating fluid pump, wherein the lubricatingfluid pump may collect lubricating fluid from the sump via a filteredinlet. In the exemplary embodiment, the lubricating fluid is oil andthus, the lubricating fluid pump is herein referred to as an oil pump.However, the term “oil pump” is used only to describe the exemplaryembodiment and other embodiments where oil is used as a lubricatingfluid, and the term shall not be construed to limit the lubricatingfluid or the lubricating fluid pump.

Referring now to FIGS. 23A and 23B, one embodiment of the engine isshown, wherein lubricating fluid is distributed to different parts ofthe engine 2200 that are located in the crankcase 2206 by a mechanicaloil pump 2208. The oil pump 2208 may include a drive gear 2210 and anidle gear 2212. In some embodiments, the mechanical oil pump 2208 may bedriven by a pump drive assembly. The pump drive assembly may include adrive shaft 2214 coupled to a drive gear 2210, wherein the drive shaft2214 includes an intermediate gear 2216 thereon. The intermediate gear2216 is preferably driven by a crankshaft gear 2220, wherein thecrankshaft gear 2220 is coupled to the primary crankshaft 2218 of theengine 2200, as shown in FIG. 24. In this configuration, the crankshaft2218 indirectly drives the mechanical oil pump 2208 via the crankshaftgear 2220, which drives the intermediate gear 2216 on the drive shaft2214, which, in turn, drives the drive gear 2210 of the oil pump 2208.

The crankshaft gear 2220 may be positioned between the crankpins 2222and 2224 of crankshaft 2218 in some embodiments, as shown in FIG. 24. Inother embodiments, the crankshaft gear 2220 may be placed at an end ofthe crankshaft 2218, as shown in FIGS. 25A-25C.

For ease of manufacturing, the crankshaft 2218 may be composed of aplurality of pieces. In these embodiments, the crankshaft gear 2220 maybe to be inserted between the crankshaft pieces during assembly of thecrankshaft.

The drive shaft 2214, in some embodiments, may be positionedperpendicularly to the crankshaft 2218, as shown in FIGS. 23A and 25A.However, in some embodiments, the drive shaft 2214 may be positionedparallel to the crankshaft 2218, as shown in FIGS. 25B and 25C.

In some embodiments, the crankshaft gear 2234 and the intermediate gear2232 may be sprockets, wherein the crankshaft gear 2234 and theintermediate gear 2232 are coupled by a chain 2226, as shown in FIGS.25C and 26C. In such an embodiments, the chain 2226 is used to drive achain drive pump (shown as 2600 in FIGS. 26A through 26C).

In some embodiments, the gear ratio between the crankshaft 2218 and thedrive shaft 2214 remains constant throughout operation. In such anembodiment, it is important to have an appropriate gear ratio betweenthe crankshaft and the drive shaft, such that the gear ratio balancesthe pump speed and the speed of the engine. This achieves a specifiedflow of lubricant required by a particular engine RPM (revolutions perminute) operating range.

In some embodiments, lubricating fluid is distributed to different partsof an engine by an electric pump. The electric pump eliminates the needfor a pump drive assembly, which is otherwise required by a mechanicaloil pump.

Referring back to FIGS. 23A and 23B, the oil pump 2208 may include aninlet 2228 to collect lubricating fluid from the sump and an outlet 2230to deliver lubricating fluid to the various parts of the engine. In someembodiments, the rotation of the drive gear 2212 and the idle gear 2210cause the lubricating fluid from the sump to be drawn into the oil pumpthrough the inlet 2228 and forced out of the pump through the outlet2230. The inlet 2228 preferably includes a filter to remove particulatesthat may be found in the lubricating fluid prior to its being drawn intothe oil pump. In some embodiments, the inlet 2228 may be connected tothe sump via a tube, pipe, or hose. In some embodiments, the inlet 2228may be in direct fluid communication with the sump.

In some embodiments, the oil pump outlet 2230 is connected to a seriesof passageways in the various engine parts, through which thelubricating fluid is delivered to the various engine parts. The outlet2230 may be integrated with the passageways so as to be in directcommunication with the passageways, or may be connected to thepassageways via a hose or tube, or a plurality of hoses or tubes. Theseries of passageways are preferably an interconnected network ofpassageways, so that the outlet 2230 may be connected to a singlepassageway inlet and still be able to deliver lubricating fluid to theengine's lubricated parts.

FIGS. 27A-27D show one embodiments, wherein the oil pump outlet (shownas 2230 in FIG. 23B) is connected to a passageway 2700 in the rockershaft 2702 of the rocking beam drive 2704. The rocker shaft passageway2700 delivers lubricating fluid to the rocker pivot bearings 2706, andis connected to and delivers lubricating fluid to the rocking beampassageways (not shown). The rocking beam passageways deliverlubricating fluid to the connecting wrist pin bearings 2708, the linkrod bearings 2710, and the link rod passageways 2712. The link rodpassageways 2712 deliver lubricating fluid to the piston rod couplingbearing 2714. The connecting rod passageway (not shown) of theconnecting rod 2720 delivers lubricating fluid to a first crank pin 2722and the crankshaft passageway 2724 of the crankshaft 2726. Thecrankshaft passageway 2724 delivers lubricating fluid to the crankshaftjournal bearings 2728, the second crank pin bearing 2730, and the splineshaft passageway 2732. The spline shaft passageway 2732 deliverslubricating fluid to the spline shaft spline joints 2734 and 2736. Theoil pump outlet (not shown, shown in FIG. 23B as 2230) in someembodiments is connected to the main feed 2740. In some embodiments, anoil pump outlet may also be connected to and provide lubricating fluidto the coupling joint linear bearings 2738. In some embodiments, an oilpump outlet may be connected to the linear bearings 2738 via a tube orhose, or plurality of tubes or hoses. Alternatively, the link rodpassageways 2712 may deliver lubricating fluid to the linear bearings2738.

Thus, the main feed 2740 delivers lubricating fluid to the journalbearings surfaces 2728. From the journal bearing surfaces 2728, thelubricating fluid is delivered to the crankshaft main passage. Thecrankshaft main passage delivers lubricating fluid to both the splineshaft passageway 2732 and the connecting rod bearing on the crank pin2724.

Lubricating fluid is delivered back to the sump, preferably by flowingout of the aforementioned bearings and into the sump. In the sump, thelubricating fluid will be collected by the oil pump and redistributedthroughout the engine.

Tube Heat Exchanger

External combustion engines, such as, for example, Stirling cycleengines, may use tube heater heads to achieve high power. FIG. 28 is across-sectional view of a cylinder and tube heater head of anillustrative Stirling cycle engine. A typical configuration of a tubeheater head 2800, as shown in FIG. 28, uses a cage of U-shaped heatertubes 2802 surrounding a combustion chamber 2804. A cylinder 2806contains a working fluid, such as, for example, helium. The workingfluid is displaced by the piston 2808 and driven through the heatertubes 2802. A burner 2810 combusts a combination of fuel and air toproduce hot combustion gases that are used to heat the working fluidthrough the heater tubes 2802 by conduction. The heater tubes 2802connect a regenerator 2812 with the cylinder 2806. The regenerator 2812may be a matrix of material having a large ratio of surface to areavolume which serves to absorb heat from the working fluid or to heat theworking fluid during the cycles of the engine. Heater tubes 2802 providea high surface area and a high heat transfer coefficient for the flow ofthe combustion gases past the heater tubes 2802. Various embodiments oftube heater heads are discussed below, and in U.S. Pat. No. 6,543,215and U.S. Pat. No. 7,308,787, which are, as previously mentioned,incorporated by reference in their entireties.

FIG. 29 is a side view in cross section of a tube heater head and acylinder. The heater head 2906 is substantially a cylinder having oneclosed end 2920 (otherwise referred to as the cylinder head) and an openend 2922. Closed end 2920 includes a plurality of U-shaped heater tubes2904 that are disposed in a burner 3036 (shown in FIG. 30). EachU-shaped tube 2904 has an outer portion 2916 (otherwise referred toherein as an “outer heater tube”) and an inner portion 2918 (otherwisereferred to herein as an “inner heater tube”). The heater tubes 2904connect the cylinder 2902 to regenerator 2910. Cylinder 2902 is disposedinside heater head 2906 and is also typically supported by the heaterhead 2906. A piston 2924 travels along the interior of cylinder 2902. Asthe piston 2924 travels toward the closed end 2920 of the heater head2906, working fluid within the cylinder 2902 is displaced and caused toflow through the heater tubes 2924 and regenerator 2910 as illustratedby arrows 2930 and 2932 in FIG. 29. A burner flange 2908 provides anattachment surface for a burner 3036 (shown in FIG. 30) and a coolerflange 2912 provides an attachment surface for a cooler (not shown).

Referring to FIG. 30, as mentioned above, the closed end of heater head3006, including the heater tubes 3004, is disposed in a burner 3036 thatincludes a combustion chamber 3038. Hot combustion gases (otherwisereferred to herein as “exhaust gases”) in combustion chamber 3038 are indirect thermal contact with heater tubes 3004 of heater head 3006.Thermal energy is transferred by conduction from the exhaust gases tothe heater tubes 3004 and from the heater tubes 3004 to the workingfluid of the engine, typically helium. Other gases, such as nitrogen,for example, or mixtures of gases, may be used, with a preferableworking fluid having high thermal conductivity and low viscosity.Non-combustible gases are used in various embodiments. Heat istransferred from the exhaust gases to the heater tubes 3004 as theexhaust gases flow around the surfaces of the heater tubes 3004. Arrows3042 show the general radial direction of flow of the exhaust gases.Arrows 3040 show the direction of flow of the exhaust gas as it exitsfrom the burner 3036. The exhaust gases exiting from the burner 3036tend to overheat the upper part of the heater tubes 3004 (near theU-bend) because the flow of the exhaust gases is greater near the upperpart of the heater tubes than at the bottom of the heater tubes (i.e.,near the bottom of the burner 3036).

The overall efficiency of an external combustion engine is dependent inpart on the efficiency of heat transfer between the combustion gases andthe working fluid of the engine.

Returning to FIG. 29, in general, the inner heater tubes 2918 are warmerthan the outer heater tubes 2916 by several hundred degrees Celsius. Theburner power and thus the amount of heating provided to the workingfluid is therefore limited by the inner heater tube 2918 temperatures.The maximum amount of heat will be transferred to the working gas if theinner and outer heater tubes are nearly the same temperature. Generally,embodiments, as described herein, either increase the heat transfer tothe outer heater tubes or decrease the rate of heat transfer to theinner heater tubes.

FIG. 31 is a perspective view of an exhaust flow concentrator and a tubeheater head in accordance with one embodiment. Heat transfer to acylinder, such as a heater-tube, in cross-flow, is generally limited toonly the upstream half of the tube. Heat transfer on the back side (ordownstream half) of the tube, however, is nearly zero due to flowseparation and recirculation. An exhaust flow concentrator 3102 may beused to improve heat transfer from the exhaust gases to the downstreamside of the outer heater tubes by directing the flow of hot exhaustgases around the downstream side (i.e. the back side) of the outerheater tubes. As shown in FIG. 31, exhaust flow concentrator 3102 is acylinder placed outside the bank of heater tubes 3104. The exhaust flowconcentrator 3102 may be fabricated from heat resistant alloys,preferably high nickel alloys such as Inconel 600, Inconel 625,Stainless Steels 310 and 316 and more preferably Hastelloy X. Openings3106 in the exhaust flow concentrator 3102 are lined up with the outerheater tubes. The openings 3106 may be any number of shapes such as aslot, round hole, oval hole, square hole etc. In FIG. 31, the openings3106 are shown as slots. In some embodiments, the slots 3106 have awidth approximately equal to the diameter of a heater tube 3104. Theexhaust flow concentrator 3102 is preferably a distance from the outerheater tubes equivalent to one to two heater tube diameters.

FIG. 32 illustrates the flow of exhaust gases using the exhaust flowconcentrator as shown in FIG. 31. As mentioned above, heat transfer isgenerally limited to the upstream side 3210 of a heater tube 3204. Usingthe exhaust flow concentrator 3202, the exhaust gas flow is forcedthrough openings 3206 as shown by arrows 3212. Accordingly, as shown inFIG. 32, the exhaust flow concentrator 3202 increases the exhaust gasflow 3212 past the downstream side 3214 of the heater tubes 3204. Theincreased exhaust gas flow past the downstream side 3214 of the heatertubes 3204 improves the heat transfer from the exhaust gases to thedownstream side 3214 of the heater tubes 3204. This in turn increasesthe efficiency of heat transfer to the working fluid which can increasethe overall efficiency and power of the engine.

Returning to FIG. 31, the exhaust flow concentrator 3102 may alsoimprove the heat transfer to the downstream side of the heater tubes3104 by radiation. Referring to FIG. 33, given enough heat transferbetween the exhaust gases and the exhaust flow concentrator, thetemperature of the exhaust flow concentrator 3302 will approach thetemperature of the exhaust gases. In a some embodiments, the exhaustflow concentrator 3302 does not carry any load and may therefore,operate at 1000.degree. C. or higher. In contrast, the heater tubes 3304generally operate at 700.degree. C. Due to the temperature difference,the exhaust flow concentrator 3302 may then radiate thermally to themuch cooler heater tubes 3304 thereby increasing the heat transfer tothe heater tubes 3304 and the working fluid of the engine. Heat transfersurfaces (or fins) 3310 may be added to the exhaust flow concentrator3302 to increase the amount of thermal energy captured by the exhaustflow concentrator 3302 that may then be transferred to the heater tubesby radiation. Fins 3310 are coupled to the exhaust flow concentrator3302 at positions outboard of and between the openings 3306 so that theexhaust gas flow is directed along the exhaust flow concentrator,thereby reducing the radiant thermal energy lost through each opening inthe exhaust flow concentrator. The fins 3310 are preferably attached tothe exhaust flow concentrator 3302 through spot welding. Alternatively,the fins 3310 may be welded or brazed to the exhaust flow concentrator3302. The fins 3310 should be fabricated from the same material as theexhaust flow concentrator 3302 to minimize differential thermalexpansion and subsequent cracking. The fins 3310 may be fabricated fromheat resistant alloys, preferably high nickel alloys such as Inconel600, Inconel 625, Stainless Steels 310 and 316 and more preferablyHastelloy X.

As mentioned above with respect to FIG. 30, the radial flow of theexhaust gases from the burner is greatest closest to the exit of theburner (i.e., the upper U-bend of the heater tubes). This is due in partto the swirl induced in the flow of the exhaust gases and the suddenexpansion as the exhaust gases exit the burner. The high exhaust gasflow rates at the top of the heater tubes creates hot spots at the topof the heater tubes and reduces the exhaust gas flow and heat transferto the lower sections of the heater tubes. Local overheating (hot spots)may result in failure of the heater tubes and thereby the failure of theengine. FIG. 34 is a perspective view of an exhaust flow axial equalizerin accordance with an embodiment. The exhaust flow axial equalizer 3420is used to improve the distribution of the exhaust gases along thelongitudinal axis of the heater tubes 3404 as the exhaust gases flowradially out of the tube heater head. (The typical radial flow of theexhaust gases is shown in FIG. 30.) As shown in FIG. 34, the exhaustflow axial equalizer 3420 is a cylinder with openings 3422. As mentionedabove, the openings 3422 may be any number of shapes such as a slot,round hole, oval hole, square hole etc. The exhaust flow axial equalizer3420 may be fabricated from heat resistant alloys, preferably highnickel alloys including Inconel 600, Inconel 625, Stainless Steels 310and 316 and more preferably Hastelloy X.

In some embodiments, the exhaust flow axial equalizer 3420 is placedoutside of the heater tubes 3404 and an exhaust flow concentrator 3402.Alternatively, the exhaust flow axial equalizer 3420 may be used byitself (i.e., without an exhaust flow concentrator 3402) and placedoutside of the heater tubes 3404 to improve the heat transfer from theexhaust gases to the heater tubes 3404. The openings 3422 of the exhaustflow axial equalizer 3420, as shown in FIG. 34, are shaped so that theyprovide a larger opening at the bottom of the heater tubes 3404. Inother words, as shown in FIG. 34, the width of the openings 3422increases from top to bottom along the longitudinal axis of the heatertubes 3404. The increased exhaust gas flow area through the openings3422 of the exhaust flow axial equalizer 3420 near the lower portions ofthe heater tubes 3404 counteracts the tendency of the exhaust gas flowto concentrate near the top of the heater tubes 3404 and therebyequalizes the axial distribution of the radial exhaust gas flow alongthe longitudinal axis of the heater tubes 3404.

In another embodiment, as shown in FIG. 35, spacing elements 3504 may beadded to an exhaust flow concentrator 3502 to reduce the spacing betweenthe heater tubes 3506. Alternatively, the spacing elements 3504 could beadded to an exhaust flow axial equalizer 3520 (shown in FIG. 34) when itis used without the exhaust flow concentrator 3504. As shown in FIG. 35,the spacing elements 3504 are placed inboard of and between theopenings. The spacers 3504 create a narrow exhaust flow channel thatforces the exhaust gas to increase its speed past the sides of heatertubes 3506. The increased speed of the combustion gas thereby increasesthe heat transfer from the combustion gases to the heater tubes 3506. Inaddition, the spacing elements may also improve the heat transfer to theheater tubes 3506 by radiation.

FIG. 36 is a cross-sectional side view, of a tube heater head 3606 andburner 3608 in accordance with an alternative embodiment. In thisembodiment, a combustion chamber of a burner 3608 is placed inside a setof heater tubes 3604 as opposed to above the set of heater tubes 3604 asshown in FIG. 30. A perforated combustion chamber liner 3615 is placedbetween the combustion chamber and the heater tubes 3604. Perforatedcombustion chamber liner 3615 protects the inner heater tubes fromdirect impingement by the flames in the combustion chamber. Like theexhaust flow axial equalizer 3420, as described above with respect toFIG. 34, the perforated combustion chamber liner 3615 equalizes theradial exhaust gas flow along the longitudinal axis of the heater tubes3604 so that the radial exhaust gas flow across the top of the heatertubes 3604 (near the U-bend) is roughly equivalent to the radial exhaustgas flow across the bottom of the heater tubes 3604. The openings in theperforated combustion chamber liner 3615 are arranged so that thecombustion gases exiting the perforated combustion chamber liner 3615pass between the inner heater tubes 3604. Diverting the combustion gasesaway from the upstream side of the inner heater tubes 3604 will reducethe inner heater tube temperature, which in turn allows for a higherburner power and a higher engine power. An exhaust flow concentrator3602 may be placed outside of the heater tubes 3604. The exhaust flowconcentrator 3602 is described above with respect to FIGS. 31 and 32.

Another method for increasing the heat transfer from the combustion gasto the heater tubes of a tube heater head so as to transfer heat, inturn, to the working fluid of the engine is shown in FIG. 37. FIG. 37 isa perspective view of a tube heater head including flow diverter fins inaccordance with an embodiment. Flow diverter fins 3702 are used todirect the exhaust gas flow around the heater tubes 3704, including thedownstream side of the heater tubes 3704, in order to increase the heattransfer from the exhaust gas to the heater tubes 3704. Flow diverterfin 3702 is thermally connected to a heater tube 3704 along the entirelength of the flow diverter fin. Therefore, in addition to directing theflow of the exhaust gas, flow diverter fins 3702 increase the surfacearea for the transfer of heat by conduction to the heater tubes 3704,and thence to the working fluid.

FIG. 38 is a top view in cross-section of a tube heater head includingflow diverter fins in accordance with an embodiment. Typically, theouter heater tubes 3806 have a large inter-tube spacing. Therefore, someembodiments as shown in FIG. 38, the flow diverter fins 3802 are used onthe outer heater tubes 3806. In an alternative embodiment, the flowdiverter fins could be placed on the inner heater tubes 3808 (also shownin FIG. 39 as 3908). As shown in FIG. 38, a pair of flow diverter finsis connected to each outer heater tube 3806. One flow diverter fin isattached to the upstream side of the heater tube and one flow diverterfin is attached to the downstream side of the heater tube. In someembodiments, the flow diverter fins 3802 are “L” shaped in cross sectionas shown in FIG. 38. Each flow diverter fin 3802 is brazed to an outerheater tube so that the inner (or upstream) flow diverter fin of oneheater tube overlaps with the outer (or downstream) flow diverter fin ofan adjacent heater tube to form a serpentine flow channel. The path ofthe exhaust gas flow caused by the flow diverter fins is shown by arrows3814. The thickness of the flow diverter fins 3802 decreases the size ofthe exhaust gas flow channel thereby increasing the speed of the exhaustgas flow. This, in turn, results in improved heat transfer to the outerheater tubes 3806. As mentioned above, with respect to FIG. 37, the flowdiverter fins 3802 also increase the surface area of the outer heatertubes 3806 for the transfer of heat by conduction to the outer heatertubes 3806.

FIG. 39 is a cross-sectional top view of a section of the tube heaterhead of FIG. 37 in accordance with an embodiment. As mentioned above,with respect to FIG. 38, a pair of flow diverter fins 3902 is brazed toeach of the outer heater tubes 3906. In some embodiments, the flowdiverter fins 3902 are attached to an outer heater tube 3906 using anickel braze along the full length of the heater tube. Alternatively,the flow diverter fins could be brazed with other high temperaturematerials, welded or joined using other techniques known in the art thatprovide a mechanical and thermal bond between the flow diverter fin andthe heater tube.

An alternative embodiment of flow diverter fins is shown in FIG. 40.FIG. 40 is a top view of a section of a tube heater head includingsingle flow diverter fins in accordance with an embodiment. In thisembodiment, a single flow diverter fin 4002 is connected to each outerheater tube 4004. In some embodiments, the flow diverter fins 4002 areattached to an outer heater tube 4004 using a nickel braze along thefull length of the heater tube. Alternatively, the flow diverter finsmay be brazed with other high temperature materials, welded or joinedusing other techniques known in the art that provide a mechanical andthermal bond between the flow diverter fin and the heater tube. Flowdiverter fins 4002 are used to direct the exhaust gas flow around theheater tubes 4004, including the downstream side of the heater tubes4004. In order to increase the heat transfer from the exhaust gas to theheater tubes 4004, flow diverter fins 4002 are thermally connected tothe heater tube 4004. Therefore, in addition to directing the flow ofexhaust gas, flow diverter fins 4002 increase the surface area for thetransfer of heat by conduction to the heater tubes 4004, and thence tothe working fluid.

FIG. 41 is a top view in cross-section of a section of a tube heaterhead including the single flow diverter fins as shown in FIG. 40 inaccordance with an embodiment. As shown in FIG. 41, a flow diverter fin4110 is placed on the upstream side of a heater tube 4106. The diverterfin 4110 is shaped so as to maintain a constant distance from thedownstream side of the heater tube 4106 and therefore improve thetransfer of heat to the heater tube 4106. In an alternative embodiment,the flow diverter fins could be placed on the inner heater tubes 4108.

Engine performance, in terms of both power and efficiency, is highest atthe highest possible temperature of the working gas in the expansionvolume of the engine. The maximum working gas temperature, however, istypically limited by the properties of the heater head. For an externalcombustion engine with a tube heater head, the maximum temperature islimited by the metallurgical properties of the heater tubes. If theheater tubes become too hot, they may soften and fail resulting inengine shut down. Alternatively, at too high of a temperature the tubeswill be severely oxidized and fail. It is, therefore, important toengine performance to control the temperature of the heater tubes. Atemperature sensing device, such as a thermocouple, may be used tomeasure the temperature of the heater tubes. The temperature sensormounting scheme may thermally bond the sensor to the heater tube andisolate the sensor from the much hotter combustion gases. The mountingscheme should be sufficiently robust to withstand the hot oxidizingenvironment of the combustion-gas and impinging flame that occur nearthe heater tubes for the life of the heater head. One set of mountingsolutions include brazing or welding thermocouples directly to theheater tubes. The thermocouples would be mounted on the part of theheater tubes exposed to the hottest combustion gas. Other possiblemounting schemes permit the replacement of the temperature sensor. Inone embodiment, the temperature sensor is in a thermowell thermallybonded to the heater tube. In another embodiment, the mounting scheme isa mount, such as a sleeve, that mechanically holds the temperaturesensor against the heater tube.

FIG. 42 is a side view in cross section of a cylinder 4204 and a burner4210. A temperature sensor 4202 is used to monitor the temperature ofthe heater tubes and provide feedback to a fuel controller (not shown)of the engine in order to maintain the heater tubes at the desiredtemperature. In some embodiments, the heater tubes are fabricated usingInconel 625 and the desired temperature is 930.degree. C. The desiredtemperature will be different for other heater tube materials. Thetemperature sensor 4202 should be placed at the hottest, and thereforethe limiting, part of the heater tubes. Generally, the hottest part ofthe heater tubes will be the upstream side of an inner heater tube 4206near the top of the heater tube. FIG. 42 shows the placement of thetemperature sensor 4202 on the upstream side of an inner heater tube4206. In some embodiments, as shown in FIG. 42, the temperature sensor4202 is clamped to the heater tube with a strip of metal 4212 that iswelded to the heater tube in order to provide good thermal contactbetween the temperature sensor 4202 and the heater tube 4206. In oneembodiment, both the heater tubes 4206 and the metal strip 4212 may beInconel 625 or other heat resistant alloys such as Inconel 600,Stainless Steels 310 and 316 and Hastelloy X. The temperature sensor4202 should be in good thermal contact with the heater tube, otherwiseit may read too high a temperature and the engine will not produce asmuch power as possible. In an alternative embodiment, the temperaturesensor sheath may be welded directly to the heater tube.

In another embodiment, as shown in FIG. 43A-B, a temperature sensormount 4320 is created with a formed strip or sheath of a refractory orhigh temperature resistant metal such as Inconel that is bonded to theexterior of the heater tube 4310. The sensor mount sheath 4320 is formedor shaped into a channel that when attached to the heater tube creates avoid that accommodates a device. In a specific embodiment, the channelis V-shaped to accommodate the insertion of a thermal sensor such as athermocouple device. The shaped channel is then bonded to the exteriorof a heater tube 4310 as shown in FIG. 43A.

FIG. 43A shows a side view of the sensor mount sheath 4320 on the heatertube 4310, while FIG. 43B is a view along the axis of the sensor mountsheath 4320. The metal should be thin enough to form, yet thick enoughto survive for the rated life of the heater head. In some embodiments,the metal is approximately between 0.005″ and 0.020″ thick. The metalmay be bent such that the bend is along the length of the strip. This“V-channel” sheath 4320 is then affixed to the exterior of the heatertube by high temperature brazing. Prior to brazing, the sheath may betack welded in several places to insure that the sheath does not moveduring the brazing process, as shown in FIG. 43A. Preferably, the brazecompound used during brazing is typically a high nickel alloy; however,any compound which will withstand the brazing temperature will work.Alternatively the sheath may be bonded to the heater tube by electronbeam or laser welding.

Now referring to FIG. 43B, a cavity 4330 is formed by affixing thesheath to the heater tube. This cavity 4330 is formed such that it mayaccept a device such as a thermocouple. When formed and brazed, thecavity may advantageously be sized to fit the thermocouple. Preferably,the fit is such that the thermocouple is pressed against the exterior ofthe heater tube. Preferably, the sheath is thermally connected to theheater tube. If the sheath is not thermally connected to the heatertube, the sheath may not be “cooled” by the working gas. The lack ofcooling may cause the sheath to operate at or near the combustion gastemperatures, which are typically high enough to eventually burn throughany metal. Brazing the sensor mount to the heater tube leads to a goodthermal contact. Alternatively, the sensor mount sheath 4320 could becontinuously welded along both sides to provide sufficient thermalconnection.

In another embodiment, as shown in FIGS. 44A-B, a second strip of metalcan be formed to create a shield 4450 over the sensor mount 4420. Theshield 4420 may be used to improve the thermal connection between thetemperature sensor, in cavity 4430, and the heater tube 4410. The shieldinsulates the sensor mount sheath 4420 from the convective heating ofthe hot combustion gases and thus improves the thermal connection to theheater tube. Furthermore, there is preferably an insulating space 4440to help further insulate the temperature sensor from the hot combustiongases as shown in FIG. 44B.

In another specific embodiment, as shown in FIGS. 45A and 45B, thetemperature sensor mount 4520 can be a small diameter tube or sleeve4540 joined to the leading edge of the heater tube 4510. FIG. 45A showsa side view of the mount on the heater tube 4510, while FIG. 45B is aview along the axis of the tube 4540 or sleeve. The sensor tube 4540 ispreferably brazed to the heater tube with a substantial braze fillet4530. The large braze fillet 4530 will maximize the thermal bond betweenthe heater tube and the sensor mount. In another embodiment, the tube orsleeve 4540 may have a shield. As described supra, an outer shield covermay help insulate the temperature sensor mount 4520 from convective heattransfer and improve the thermal connection to the heater tube.

In an alternative embodiment of the tube heater head, the U-shapedheater tubes may be replaced with several helical wound heater tubes.Typically, fewer helical shaped heater tubes are required to achievesimilar heat transfer between the exhaust gases and the working fluid.Reducing the number of heater tubes reduces the material and fabricationcosts of the heater head. In general, a helical heater tube does notrequire the additional fabrication steps of forming and attaching fins.In addition, a helical heater tube provides fewer joints that couldfail, thus increasing the reliability of the heater head.

FIGS. 46A-46D are perspective views of a helical heater tube inaccordance some embodiments. The helical heater tube, 4602, as shown inFIG. 46A, may be formed from a single long piece of tubing by wrappingthe tubing around a mandrel to form a tight helical coil 4604. The tubeis then bent around at a right angle to create a straight return passageout of the helix 4606. The right angle may be formed before the finalhelical loop is formed so that the return can be clocked to the correctangle. FIGS. 46B and 46C show further views of the helical heater tube.FIG. 46D shows an alternative embodiment of the helical heater tube inwhich the straight return passage 4606 goes through the center of thehelical coil 4604. FIG. 47 shows a helical heater tube in accordancewith one embodiment. In FIG. 47, the helical heater tube 4702 is shapedas a double helix. The heater tube 4702 may be formed using a U-shapedtube wound to form a double helix.

FIG. 48 is a perspective view of a tube heater head with helical heatertubes (as shown in FIG. 46A) in accordance with one embodiment. Helicalheater tubes 4802 are mounted in a circular pattern to the top of aheater head 4803 to form a combustion chamber 4806 in the center of thehelical heater tubes 4802. The helical heater tubes 4802 provide asignificant amount of heat exchange surface around the outside of thecombustion chamber 4806.

FIG. 49 is a cross sectional view of a burner and a tube heater headwith helical heater tubes in accordance with some embodiments. Helicalheater tubes 4902 connect the hot end of a regenerator 4904 to acylinder 4905. The helical heater tubes 4902 are arranged to form acombustion chamber 4906 (also shown in FIG. 50 as 5006) for a burner4907 that is mounted coaxially and above the helical heater tubes 4902.Fuel and air are mixed in a throat 4908 of the burner 4907 and combustedin the combustion chamber 4906. The hot combustion (or exhaust) gasesflow, as shown by arrows 4914, across the helical heater tubes 4902,providing heat to the working fluid as it passes through the helicalheater tubes 4902.

In one embodiment, the heater head 4903 (also shown in FIG. 50 as 5003)further includes a heater tube cap 4910 at the top of each helicalcoiled heater tubes 4902 to prevent the exhaust gas from entering thehelical coil portion 4901 (also shown in FIG. 50 as 5001) of each heatertube and exiting out the top of the coil. In another embodiment, anannular shaped piece of metal covers the top of all of the helicalcoiled heater tubes. The heater tube cap 4910 prevents the flow of theexhaust gas along the heater head axis to the top of the helical heatertubes between the helical heater tubes. In one embodiment, the heatertube cap 4910 may be Inconel 625 or other heat resistant alloys such asInconel 600, Stainless Steels 310 and 316 and Hastelloy X.

In another embodiment, the top of the heater head 4903 under the helicalheater tubes 4902 is covered with a moldable ceramic paste. The ceramicpaste insulates the heater head 4903 from impingement heating by theflames in the combustion chamber 4906 as well as from the exhaust gases.In addition, the ceramic blocks the flow of the exhaust gases along theheater head axis to the bottom of the helical heater tubes 4902 eitherbetween the helical heater tubes 4902 or inside the helical coil portion4901 of each heater tube.

FIG. 50 is a top view of a tube heater head with helical heater tubes inaccordance with one embodiment. As shown in FIG. 50, the return orstraight section 5002 of each helical heater tube 5000 is advantageouslyplaced outboard of gap 5009 between adjacent helical heater tubes 5000.It is important to balance the flow of exhaust gases through the helicalheater tubes 5000 with the flow of exhaust gases through the gaps 5009between the helical heater tubes 5000. By placing the straight portion5002 of the helical heater tube outboard of the gap 5009, the pressuredrop for exhaust gas passing through the helical heater tubes isincreased, thereby forcing more of the exhaust gas through the helicalcoils where the heat transfer and heat exchange area are high. Exhaustgas that does not pass between the helical heater tubes will impinge onthe straight section 5002 of the helical heater tube, providing highheat transfer between the exhaust gases and the straight section. BothFIGS. 49 and 50 show the helical heater tubes placed as close togetheras possible to minimize the flow of exhaust gas between the helicalheater tubes and thus maximize heat transfer. In one embodiment, thehelical coiled heater tubes 4901 may be arranged so that the coils nesttogether.

Pin or Fin Heat Exchanger

Now referring to FIGS. 51A and 51B, fins or pins may alternatively beused to increase the interfacial area between the hot fluid combustionproducts and the solid heater head so as to transfer heat, in turn, tothe working fluid of the engine. Heater head 5100 may have heat transferpins 5124, here shown on the interior surface of heater head 5100, inthe space between the heater head and expansion cylinder liner 5115.Additionally, as shown in FIG. 51B in a cross section of Stirling cycleengine 5196 taken along a different diameter of expansion volume 5198from that of FIG. 51A, heat transfer pins 5130 may also be disposed onthe exterior surface of heater head 5100 so as to provide a largesurface area for the transfer of heat by conduction to heater head 5100,and thence to the working fluid, from combustion gases flowing fromcombustor 5122 past the heat transfer pins. Dashed line 5131 representsthe longitudinal axis of the expansion cylinder. FIG. 51B also showsheat transfer pins 5133 lining the interior and exterior surfaces of thetop of heater head 5100, in accordance with one embodiment.Interior-facing heat transfer pins 5124 serve to provide a large surfacearea for the transfer of heat by conduction from heater head 5100 toworking fluid displaced from expansion volume 5198 by the expansionpiston and driven through regenerator chamber 5132. Additionalembodiments of heater head 5100 are disclosed in U.S. Pat. No.6,381,958, and No. 6,966,182, which, as previously mentioned, areincorporated by reference in their entireties.

Depending on the size of heater head 5100, hundreds or thousands ofinner transfer pins 5124 and outer heat transfer pins 5130 may bedesirable.

One method for manufacturing heater head 5100 with heat transfer pins5124 and 5130 includes casting the heater head and pins (or otherprotuberances) as an integral unit. Casting methods for fabricating theheater head and pins as an integral unit include, for example,investment casting, sand casting, or die casting.

While the use of pin fins is known for improving heat transfer between asurface and a fluid, the integral casting of radial pin fins on thecylindrical heater head of a Stirling engine has not been practiced norsuggested in the art, despite the fact that casting the heater head andit's heat exchange surfaces in a single step is one of the most costeffective methods to produce a heater head. The difficulty encounteredin integral casting of radial pin fins is discussed further below. A pinfin that could be cast as part of cylindrical wall would allow theinexpensive fabrication of a highly effective heater head and/or coolerfor a Stirling engine.

Castings are made by creating negative forms of the desired part. Allforms of production casting (sand, investment and injection) involvesforming extended surfaces and details by injecting material into a moldand then removing the mold from the material leaving the desirednegative or positive form behind. Removing the mold from the materialrequires that all the extended surfaces are at least parallel. In fact,good design practice requires slight draft on these extended surfaces sothat they release cleanly. Forming radial pins on the outside or insideof a cylinder would require the molds to contain tens or hundreds ofparts that pull apart in different directions. Such a mold would be costprohibitive.

In accordance various embodiments, pins or fins may be cast onto theinside and outside surface of Stirling heat exchangers using productionsand, investment or metal injection casting methods. Referring to FIGS.52A-52D and 53D, and, first, to FIG. 52A, pins 5202 are arranged intoseveral groups 5208 of parallel pins 5202 around cylindrical wall 5210of heater head 5100, shown in cross section parallel to the central axisin FIG. 52B and in cross section transverse to the central axis, in FIG.52C. It should be noted that the technology herein described mayadvantageously be applied more generally in any other heat exchangerapplication. All the pins 5202 in each group 5208 are parallel to eachother. Only the pins 5202 in the center of the group are truly radial.The pins on the outside of the group, such as those designated bynumeral 5204 in FIGS. 52C and 53D, are angled inward from a local radiussuch as to be substantially parallel to a radial line 5212 toward thecenter of the group. In addition, the pins on the outside of the groupare preferably longer, typically by a small amount, than pins closer tothe center of the group. However, the heat transfer only changes onlyslightly from the center of the group to the outside in the embodimentdepicted in FIGS. 52A-52C, and 53D in which 5 groups 5208 of parallelpins provide approximately radial pin fins around cylinder 5210.

In the casting process in accordance with some embodiments, positive ornegative molds of each group of parallel fins are formed in a singlepiece. Several mold pieces are then assembled to form the negative formfor a sand casting. In investment mold casting, the wax positive can beformed in an injection mold with only a handful of separate parts thatpull apart in different directions. The resulting mold is formed at anacceptable cost, thereby making production of a pin fin heater headeconomically practical.

Casting of a heater head having protuberances, such as pins, extendingto the interior and exterior of a part with cylindrical walls may beachieved, in accordance with various embodiments, by investment, orlost-wax, casting, as well as by sand casting, die casting, or othercasting processes. The interior or exterior protuberances, or both, maybe integrally cast as part of the head.

While typically more cheaply accomplished than machining or assembly ofthe pin arrays, casting pin arrays may still have attendant difficultiesand substantial costs. Additionally, the casting process may result in aheater head that is less than fully densely populated with pins, thusincreasing the fraction of gases failing to collide with the heater headsurface and reducing the efficiency of heat transfer.

One embodiment of the method for populating the surfaces of heater head5100 with heat transfer pins entails fabrication of heater 5100 andarrays of heat transfer pins in separate fabrication processes. An array5250 (also shown in FIG. 53B as 5350) of heat transfer pins 5252 may becast or injection molded with panel 5254 resulting in an integralbacking panel structure shown in FIG. 52D. Pin arrays 5250, aftercasting or molding, are mounted to the inner and outer surfaces of theheater head by a high temperature braze. Thus, a more densely populatedhead with a resultant low rate of gas leakage past the pins mayadvantageously be achieved. In other embodiments, panels 5254 may besecured by various mechanical means to the heater head.

Transient liquid-phase (TLP) bonding, as described, for example, in theAerospace Structural Metals Handbook, Code 4218, p. 6 (1999) isparticularly advantageous for brazing the panels to the head, sincenickel based superalloys, typically employed for fabrication of thehead, is difficult to weld by conventional processes, and operates in ahigh stress and high temperature environment. Advantages of TLP bondingin this application are that the parts braced by TLP are effectivelywelded using the parent material and have nearly the same tensilestrength properties as integrally cast parts. TLP bonds do not remelt atelevated temperatures, whereas typical brazes will remelt at the brazingtemperature. This is of particular significance in the case ofcontinuous operation at elevated temperatures where temperatureexcursions may occur, as in the present application.

The panels 5254 of pins may be attached to the interior or exterior ofeither the heater head or the cooler by other means. In one alternativeembodiment, the panel may be mechanically attached into slots at itslateral edges. The slots are provided in dividers 5306 (described in thefollowing discussion). In another embodiment, the panels are attached tothe heater head or cooler by brazing. In yet another embodiment, thepanels are attached to the heater head or cooler by sintering the panelsto the cylindrical walls of the heater head or cooler.

Dividers 5306, as shown in FIGS. 52C, 53A, and 53B, may advantageouslyimprove the heat transfer rate of the pin fin panels. Additionally, theymay provide a convenient location for locating temperature sensors.Lastly, the dividers may advantageously provide a convenient structureto which to attach panels of pins to the heater head, in one embodiment,and a parting line for casting operations, in accordance with a furtherembodiment.

Dividers 5306 may serve to improve the thermal effectiveness of the pinfin arrays in the following manner. Referring, once again, to FIG. 52A,the rate of heat transfer for a fluid flowing through staggered pin finsis significantly higher than for fluid flowing through aligned pin fins.Fluid approaching a staggered pin array 5208 would travel at a 45-degreeangle to an axial path along the length of the cylinder, with the skewdirection designated by numeral 5214. In order to provide for improvedthermal transfer, dividers 5206, 5306 are provided, in accordance someembodiments, to force the fluid flow through the staggered array of pinfins along a path designated by numeral 5212. In addition to forcing theflow to travel axially, the dividers provide convenient interfaces andjoining planes for the casting molds described above.

In certain embodiments, individual arrays 5250, each with its associatedpanel segment 5254, comprise arcuate fractions of the circumferentialdistance around the heater head. This is apparent in the top view of theheater head assembly shown in perspective in FIG. 53A. Cylinder head5320 is shown, as is exterior surface 5302 of the heater head. Backersegments supporting arrays of heat transfer pins are not shown but areinserted, during assembly, in spaces 5304 surrounding exterior surface5302 of the heater head. Between successive heat transfer pin arraysegments are trapezoidal dividers 5306 which are baffled to block theflow of exhaust gases in a downward direction through any path otherthan past the heat transfer pins.

In one embodiment, flow dividers 5306 include structures formechanically retaining the panel segments 5254 during assembly, beforebrazing, or simply to mechanically retain the panels 5254 against heaterhead 5302.

In order to maximize engine power, the hottest part of the heater headis preferably at the highest temperature allowed, considering themetallurgical creep and tensile strength, stress, and appropriatefactors of safety. Maintaining the hottest part of the heater head atthe highest temperature requires measuring the temperature of thehottest part of the heater head. The dividers provide a convenientlocation and routing for temperature sensors on the heater had to anyaxial location along the pin fin arrays. Hot gas flow path 5313 (shownalso in FIG. 51A), is defined, on the outside, by gas flow channel cover5340. Since exhaust gases do not flow through dividers 5306, atemperature sensor, such as thermocouple 5138 (shown in FIGS. 51A and53C) is advantageously disposed in divider 5306 in order to monitor thetemperature of heater head 5100 with which the temperature sensor is inthermal contact. The position of pin arrays 5250 and temperature sensor5138 mounted within divider 5306 is shown more clearly in the view ofFIG. 53B in which the pin backer has been removed.

Temperature sensing device 5138 is preferably disposed within divider5306 as depicted in FIG. 53B. More particularly, temperature sensing tip5339 of temperature sensor 5138 is preferably located in the slotcorresponding to divider 5306 as nearly as possible to cylinder head5320 in that this area is typically the hottest part of the heater head.Alternatively, temperature sensor 5138 may be mounted directly tocylinder head 5320, however location of the sensor in the slot, asdescribed, is used in some embodiments. Engine performance, in terms ofboth power and efficiency, is highest at the highest possibletemperature, yet the maximum temperature is typically limited bymetallurgical properties. Therefore, sensor 5138 should be placed tomeasure the temperature of the hottest, and therefore the limiting, partof the heater head. Additionally, temperature sensor 5138 should beinsulated from combustion gases and walls of divider 5306 by ceramicinsulation 5342, as shown in FIG. 53C. The ceramic can also form anadhesive bond with the walls of the divider to retain the temperaturesensor in place. Electrical leads 5344 of temperature sensor 5138 shouldalso be electrically insulated.

Although the burner is designed to have circumferential symmetry, hotspots may develop on heater head 5320. Adding to the problem, the alloystypically employed for fabrication of the heater head, on account oftheir high melting point, have relatively poor thermal conductivity.Once hot spots form, they are apt to endure because the gas flow outsidethe head is axial rather than circumferential, since dividers 5306(shown in FIG. 53A) impede any circumferential flow. Additionally,heating may increase local gas viscosity thereby redirecting more flowto other channels. In order to even out the temperature distribution onthe heater head, a layer of highly thermally conductive metal, such ascopper, of thickness greater than 0.001 in. and preferably about 0.005in. is applied to interior surface 5348 of heater head 5320, bydeposition or plating, or other application method. Alternatively, asimilar coating may be applied to the exterior surface, in accordancewith another embodiment.

In order to keep the size of the Stirling cycle engine small, it isimportant to maximize the heat flux from the combustion gas through theheater head. Whereas prior art employed loops of pipe in which heattransfer to the working fluid is achieved, loops engender both lowreliability (since the loops are mechanically vulnerable) and highercost, due to the more complicated loop geometry and extra materials. Thelimiting constraint on the heat flux are the thermo-mechanicalproperties of the heater head material that must be able to withstandthe high temperatures of the combustion chamber while maintaining thestructural integrity of the pressurized head. The maximum designtemperature is determined by the hottest point on the heater head whichis typically at the top of the wall. Ideally, the entire heater wall hotsection would be at this maximum temperature, as may be controlled, forexample, by controlling the fuel flow.

As combustion gases travel past the heater head in gas flow channels5113, 5313 (shown in FIG. 51A), the gas temperature decreases as heat istransferred from the gas to the heater head. As a result, the maximumallowed heater head temperature at the top of the gas flow channel mustbe set by the material used for the heater head. The material ispreferably chosen from the family of high nickel alloys, commonly knownas super alloys, such as Inconel 600 (having a maximum temperatureT.sub.max=800.degree. C. before softening), Inconel 625(T.sub.max=900.degree. C.), Inconel 754 (T.sub.max=1080.degree. C.), orHastelloy GMR 235 (T.sub.max=935.degree. C.). The gas in gas channel5113, 5313 may cool by as much as 350.degree. C. on transit through thechannel, resulting in underheating of the bottom of the hot zone.

In accordance with some embodiments, the temperature profile of theheater wall is controlled by means of heat transfer geometry, as nowdescribed. One method for controlling the geometry is by means ofproviding a variable cross-section gas flow channel 5113, 5313 (shown inFIGS. 51A and 54A). The radial dimension (perpendicular to the wall ofthe heater head), and thus the cross-section of the channel, is large atthe top of the heater wall, thereby allowing much of the gas to bypassthe pin array at the top of the wall. The bypass allows hotter gas toreach the pin array at the bottom of the wall thereby allowing thebottom pin array to operate closer to its maximum temperature. Thetemperature gradient from the top of the heater to the bottom of the hotsection (before regenerator volume 5132, shown in FIG. 51A) has beenreduced from as much as 350.degree. C. to 100.degree. C. using avariable cross-section gas flow channel.

A second method for controlling the geometry is by varying thepopulation density and the geometry of the pin array as a function ofposition along the gas flow channel. The geometry of the pins may beadjusted by varying the height/diameter (H/D) ratio of the pins. If acasting process is used to form the pin array, the range of H/D rationsmay be limited by the process. If pin rings are used, the range of H/Dratios may be extended.

Referring now to FIGS. 53E, 53F, 54A and 54B, arrow 5402 designates thepath of heated exhaust gases past heater head 5100. Outer heat transferpins 5130 intercept the heated exhaust gases and transfer heat viaheater head 5100 and inner heat transfer pins 5124 to the working fluidthat is driven from expansion cylinder 5115 along path 5404. (Forclarity, heat transfer pins 5130 and 5124 are shown schematically inFIG. 54A. Additional heat transfer pins 5130 and 5124 had been depicted,not to scale, in the view of FIGS. 53E, 53F, and 54B.) Successive heattransfer pins 5406, 5408, and 5410, for example, present a progressivelylarger cross section to the flow of exhaust gas along path 5402. Thus,while the exhaust gas has transferred some fraction of its heat prior toarrival at the lower pins, heat is extracted there with a greaterconduction rate, thereby reducing the temperature gradient between thetop 5412 and bottom 5414 of the path of working fluid between expansionvolume 5198 and regenerator volume 5132. Typical temperatures of thesurface of expansion cylinder 5115 are indicated in FIG. 54A:850.degree. C. at the top of the cylinder, 750.degree. C. at the centerof the cylinder, and 600.degree. C. at the end of the cylinder closestto the regenerator volume.

Another method for achieving more even distribution of heat from theexhaust gases to the heater head is to create a tapered divider on theoutside diameter of the heater head by means of concentric tapered pinbacker 5146, as shown in FIG. 54A. The cross-sectional view of FIG. 54Ashows how tapered pin backer 5146 allows some of the hottest exhaust gasto bypass the pins near the top of the heater head. Pin backer 5146creates a narrowing annular gap on the outside of the pins thatprogressively forces more and more of the exhaust gases into the pinheat exchanger.

Another method for increasing the surface area of the interface betweena solid such as heater head 5100 and a fluid such as combustion gases asdiscussed above is now described with reference to FIGS. 55A-55D. Aneffect analogous to that of fabricating heat transfer pins by casting orotherwise may be obtained by punching holes 5160 into a thin annularring 5162 shown in top view in FIG. 55A and in side view in FIG. 55B.The thickness of ring 5162, which may be referred to as a ‘heat transferpin ring’ is comparable to the thickness of the heat transfer pinsdiscussed above, and is governed by the strength of the heat-conductivematerial at the high temperature of the combustion gases traversingholes 5160. The shape and disposition of holes 5160 within each ring isa matter of design for a particular application, indeed, holes 5160 maynot be surrounded by solid material. The material of rings 5162 ispreferably an oxidation-resistant metal such as Inconel 625 or HastelloyGMR 235, though other heat-conducting materials may be used. Rings 5162may be produced inexpensively by a metal stamping process. Rings 5162are then mounted and brazed, or otherwise bonded, to the outer surfaceheater head 5100, as shown with respect to outer pin rings 5164 in FIG.55C, and with respect to inner pin rings 5166 in FIG. 55D. Additionalrings may be interspersed between the pin rings to control the verticalspacing between the pins. Expansion cylinder liner 5115 is shown in theinterior of inner pin rings 5166.

Heat transfer rings 5162 may be advantageously applied to the interiorof the heater head as well as to both the exterior and interior of thecooler of a thermal cycle engine. In these applications, the rings neednot be oxidation resistant. Materials including copper and nickel arepreferably used on the interior of the heater head, while the rings forthe cooler are preferably made of one of various high thermalconductivity materials including aluminum, copper, zinc, etc.

The total cross sectional area of the heat transfer pins taken in aslice perpendicular to cylinder axis 5168 need not be constant, indeed,it is advantageously varied, as discussed in detail above, in referenceto FIG. 54.

Referring to FIGS. 56A through 56C, the interior or exterior heatexchange surfaces may also be formed from various folded fin structures5600, 5602, or 5604. The folded fin structures may be made of materialsimilar to that of the heater head pressure dome or of high thermalconductivity materials such as copper which may provide improved finefficiency. Fins fabricated from high melting-point materials such asthat of the heater head 5100 (shown in FIG. 51A) may be continuous fromthe top to the bottom of the heater head. Folded fins may be fabricatedfrom sheet metal and brazed to the interior surface of the heater head.Three folded fin configurations are shown by way of example: wavy fins5600, lanced fins 5602, and offset fins 5604. In each case, the gas flowdirection is indicated by an arrow designated by numeral 5606.

Fins formed from a dissimilar metal to that of heater head 5100 areattached in axial segments to avoid differential thermal expansion frombreaking the brazed joint between the fins and the head. The offset finconfiguration of FIG. 56C advantageously provides a superior heattransfer coefficient to that of plain fins.

The use of high thermal conductivity metal for the folded fins mayadvantageously allow the fins to be made longer, thereby improving heattransfer and reducing resistance to flow of the gas and improving engineefficiency.

Heater Head Support Ribs

The walls of the heater head must be sufficiently strong, at operatingtemperatures, to withstand the elevated pressure of the working gas. Itis typically desirable to operate Stirling cycle engines at as high aworking gas pressure as possible, thus, enabling the head to withstandhigher pressures is highly advantageous. In designing the heater head,it must be borne in mind that increasing the pressure at a givenoperating temperature typically requires increasing the heater head wallthickness in direct proportion. On the other had, thickening the heaterhead wall results in a longer thermal conduction path between theexterior heat source and the working gas.

Moreover, thermal conduction increases with heat exchanger surface area,thus thermal efficiency is increased by increasing the diameter of theheater head. Stress in the wall, however, is substantially proportionalto the diameter of the head, thus increasing the head diameter, at agiven temperature and interior gas pressure, requires increasing thewall thickness in direct proportion.

The strength considerations are tantamount at typical Stirling enginehead temperatures, in fact, they drive the maximum operatingtemperature, since, as discussed, efficiency increases with temperature.Both creep and ultimate tensile strengths of materials tend to fall offprecipitously when specified elevated temperatures are reached.Referring to FIG. 57A, the yield strength at 0.2% offset and ultimatetensile strength are shown for the GMR 235 nickel alloy in typicalrepresentation of the qualitative behavior of nickel alloys. Similarly,in FIG. 57B, it can be seen that the 0.01% per hour creep rate strengthof GMR 235 falls from 40 ksi to half as the temperature rises from1500.degree. F. to 1700.degree. F.

Some embodiments provide interior ribs (or hoops) 5800, such as thosedisclosed in U.S. Pat. No. 6,381,958, and No. 6,966,182, that enhancestructural support of heater head 5801, as shown in cross-section inFIG. 58. Ribs 5800 are characterized by an interior bore 5802. The creepstrength and rupture strength of heater head 5801 is thus determinedpredominantly by an effective thickness 5804 of the heater head and theinterior bore diameter 5802. Heat conduction through the heater head isnot limited by thickness 5804 since intervening segments 5806 of thehead are narrower and provide enhanced heat conduction. Ribs 5800 notonly relieve hoop stresses on outer wall 5808 of head 5801 butadditionally provide supplemental surface area interior to the heaterhead and thus advantageously enhance heat transfer to the working fluid.

Further advantages of providing ribs 5800 interior to the heater headinclude reducing the temperature gradient across the head wall 5808 fora given rate of heat transfer, as well as allowing operation at higherhot end working temperatures. Additionally, by reducing the stressrequirements on the outer wall, alternative materials to nickel basedsuperalloys may be used, advantageously providing superior conductivityat reduced cost.

A cross section of heater head 5801 with ribs 5800 is further shown inFIG. 59. Dashed line 5910 designates the central longitudinal axis ofthe expansion cylinder. In accordance with various embodiments expansioncylinder hot sleeve 5912 may have transverse flow diverters 5914 fordirecting the flow of working gas, represented by around 5916, aroundcircumferential ribs 5800 for enhancing heat transfer to the workinggas. The additional width h of ribs 5800 contributes to the hoopstrength of heater head 5101, whereas heat transfer is governedpredominantly by the narrower thickness t of outer heater head wall5808. In typical Stirling engine applications, while the heater headexterior may be run as hot as 1800.degree. F., ribs 5800 that providestructure strength typically run no hotter than 1300.degree. F.

Advantages of enhanced hoop strength concurrent with enhanced thermalconductivity, as discussed above with reference to FIG. 58 mayadditionally be obtained in accordance with several alternateembodiments. Referring to FIGS. 60A and 60B, cross sections are shown ofa heater head 6030, wherein tubular openings 6032 run parallel to heaterhead wall 6008. As shown in the cross sectional view of FIG. 60B, takenalong line AA, tubes 6032 allow working gas to pass down the wall,enhancing heat transfer from outside the head to the working gas.Additionally, the wall 6008 may be thicker, for the same rate of heattransfer, thus providing additional strength. Moreover, the thick wallsection 6010 (also shown in FIG. 61B as 6110) interior to passages 6032remains cooler than would otherwise be the case, providing furtheradditional strength. Heater head 6030 is preferably cast with tubularpassages 6032 which may be round in cross section or of other shapes.

FIG. 61A shows a further heater head 6140 wherein tubular openings 6132run parallel to heater head wall 6108 and are interrupted by openingsthat run out to thinner sections 6142 of the heater head wall. As shownin the cross sectional view of FIG. 62B, taken along line AA, tubes 6132allow working gas to pass down the wall, enhancing heat transfer fromoutside the head to the working gas to a degree substantially enhancedover that of the straight tube design shown in FIGS. 62A and 62B.Additionally, openings 6144 provide additional area for removal ofceramic cores used in the casting process to create such long, thinholes. Increased access to the holes allows faster chemical leaching ofthe core in the course of the manufacturing process.

FIG. 62B shows yet another heater head 6250, wherein ribs 6252 aredisposed in a helix within heater head wall 6208, thereby providing thewall with enhanced rigidity in both the circumferential and axialdirections. The working gas flows through the spiral 6254 on a pathbetween the expansion piston and the heater head, on its way to theregenerator. FIG. 62B shows a transverse cross section of the heaterhead of FIG. 62A taken along line AA. Various embodiments includeemploying a linear, or other, approximation to spiral 6254, to obtaincomparable advantages of stiffening and heat transfer.

Heater head 6250 of FIGS. 62A and 62B is preferably fabricated bycasting. A side view of core assembly 6260 for use in the castingprocess is shown in FIG. 62C. It is additionally advantageous to provideribs for internal support of the dome of the heater head and to provideadditional heat exchange on the dome, thereby cooling the inner surfaceof the dome. The complementary core structure of the dome is shown inFIG. 62D, and, in cross section, as viewed from the top, in FIG. 62D. Aperspective view of core assembly 6260 is shown in FIG. 62E.

It is to be understood that the various heater head embodiments andmethods for their manufacture described herein may be adapted tofunction in a multiple heater head configuration.

Regenerator

A regenerator is used in a Stirling cycle machine, as discussed aboveand as described in U.S. Pat. No. 6,591,609, and No. 6,862,883, to addand remove heat from the working fluid during different phases of theStirling cycle. The regenerator used in a Stirling cycle machine must becapable of high heat transfer rates which typically suggests a high heattransfer area and low flow resistance to the working fluid. Low flowresistance also contributes to the overall efficiency of the engine byreducing the energy required to pump the working fluid. Additionally, aregenerator must be fabricated in such a manner as to resist spalling orfragmentation because fragments may be entrained in the working fluidand transported to the compression or expansion cylinders and result indamage to the piston seals.

One regenerator design uses several hundred stacked metal screens. Whileexhibiting a high heat transfer surface, low flow resistance and lowspalling, metal screens may suffer the disadvantage that their cuttingand handling may generate small metal fragments that must be removedbefore assembling the regenerator. Additionally, stainless steel wovenwire mesh contributes appreciably to the cost of the Stirling cycleengine.

A three dimensional random fiber network, such as stainless steel woolor ceramic fiber, for example, may be used as the regenerator, as nowdescribed with reference to FIG. 63A. Stainless steel wool regenerator6300 advantageously provides a large surface area to volume ratio,thereby providing favorable heat transfer rates at low fluid flowfriction in a compact form. Additionally, cumbersome manufacturing stepsof cutting, cleaning and assembling large numbers of screens areadvantageously eliminated. The low mechanical strength of steel wool andthe tendency of steel wool to spall may both be overcome as nowdescribed. In some embodiments, the individual steel wires 6302 and 6304are “cross-linked” into a unitary 3D wire matrix.

The starting material for the regenerator may be fibrilose and of randomfiber form such as either steel or nickel wool. The composition of thefiber may be a glass or a ceramic or a metal such as steel, copper, orother high temperature materials. The diameter of the fiber ispreferably in the range from 10 micrometers to 1 millimeter depending onthe size of the regenerator and the properties of the metal. Thestarting material is placed into a form corresponding to the final shapeof the regenerator which is depicted in cross-section in FIG. 63B. Innercanister cylindrical wall 6320, outer canister cylindrical wall 6322,and regenerator network 6300 are shown. The density of the regeneratoris controlled by the amount of starting material placed in the form. Theform may be porous to allow fluids to pass through the form.

In some embodiments, unsintered steel wool is employed as regeneratornetwork 6300. Regenerator network 6300 is then retained within theregenerator canister by regenerator retaining screens 6324 or otherfilter, thereby comprising a “basket” which may advantageously capturesteel wool fragments.

In one embodiment, applicable to starting material that is electricallyconducting, the starting material is placed in a porous form and placedin an electrolyte bath. The starting material may be a metal, such asstainless steel, for example. An electrical connection is made with thestarting material thereby forming an electrode. Cross-linking of theindividual fibers in the starting material is accomplished byelectrically depositing a second material 6306 onto the startingmaterial. The selection of the starting material will depend on suchfactors as the particular deposition technique chosen and the chemicalcompatibility of the first and second materials, as known to one ofordinary skill in the electrochemical art. During deposition, the secondmaterial will build up on the starting material and form bridges 6308between the individual fibers of the starting material in places wherethe individual fibers are in close proximity to each other. Thedeposition is continued until the bridges have grown to a sufficientsize to hold the two individual fibers rigidly in place.

The deposition duration depends on the particular deposition process andis easily determined by one of ordinary skill in the art. After thedeposition is completed, the regenerator is removed from the bath andthe form and is cleaned.

In another embodiment the starting material is placed in a form that maybe porous or not. The form containing the starting material is placed ina furnace and is partially sintered into a unitary piece. The selectionof the sintering temperature and sintering time is easily determined byone of ordinary skill in the sintering art.

In another embodiment the starting material is placed in a porous form.The form containing the starting material is placed in a chemical bathand a second material, such as nickel, is chemically deposited to formbridges between the individual fibers.

In another embodiment the starting material is a silica glass fiberwhich is placed into a porous form. The glass fiber and form is dippedin a solution of tetraethylorthosilicate (TEOS) and ethanol so that thefiber is completely wetted by the solution. The fiber and form areremoved from the solution and allowed to drain in a humid atmosphere.The solution will form meniscoidal shapes bridging fibers in closeproximity to each other. The humidity of the atmosphere will start thehydrolysis-condensation reaction that converts the TEOS to silicaforming a cross link between the two fibers. The fiber and form may beheat treated at a temperature less than 1000° C., most preferably lessthan 600° C., to remove the reactant products and form a silica bridgebetween the fibers.

In another embodiment a ceramic slurry is deposited onto a reticulatedfoam having the shape of the regenerator. The slurry is dried on thereticulated foam and heat treated to burn off the foam and sinter theceramic. The ceramic may be composed of an oxide ceramic such ascordierite, alumina, or zirconia. The composition of the ceramic slurryand the heat treatment profile is easily specified by one of ordinaryskill in the ceramic processing art.

In yet other embodiments, knit or woven wire is employed in fabricationof a regenerator as now described with reference to FIG. 64A. Inaccordance with these embodiments, knit or woven wire tube 6401 isflattened by rollers 6402 into tape 6404, in which form it is woundabout mandrel 6406 into annular layers 6408. Stainless steel isadvantageously used for knit wire tube 6401 because of its ability towithstand elevated temperature operation, and the diameter of the wireused is typically in the range of 1-2 mils, however other materials andgauges may be used in various embodiments. Alternatively, a plurality,typically 5-10, of the stainless steel wires may be loosely wound into amulti-filament thread prior to knitting into a wire tube. This processadvantageously strengthens the resulting tube 6401. When mandrel 6406 isremoved, annular assembly 6410 may be used as a regenerator in a thermalcycle engine.

Still another embodiment is now described with reference to FIGS. 64Bthrough 64E. Knit or woven wire tube 6401, shown in its rightcylindrical form in FIG. 64B, is shown scored and partially compressedin FIG. 64C. Alternatively, the scoring may be at an angle 6414 withrespect to the central axis 6412 of the tube, as shown in FIG. 64D. Tube6401 is then axially compressed along central axis 6412 to form thebellows form 6416 shown in FIG. 64E that is then disposed as aregenerator within the regenerator volume 408 (shown in FIG. 4) of aStirling cycle engine.

It is to be understood that the various regenerator embodiments andmethods for their manufacture described herein may be adapted tofunction in a multiple cylinder configuration.

Coolant Penetrating Cold-End Pressure Vessel

Referring now to FIGS. 65A-C, various cross-sections of an engine, suchas a Stirling cycle engine, are shown in accordance with someembodiments. Engine 6500 is hermetically sealed. A crankcase 6502 servesas the cold-end pressure vessel and contains a charge gas in an interiorvolume 6504. Crankcase 6502 can be made arbitrarily strong withoutsacrificing thermal performance by using sufficiently thick steel orother structural material. A heater head 6506 serves as the hot-endpressure vessel and is preferably fabricated from a high temperaturesuper-alloy such as Inconel 625, GMR-235, etc. Heater head 6506 is usedto transfer thermal energy by conduction from an external thermal source(not shown) to the working fluid. Thermal energy may be provided fromvarious heat sources such as solar radiation or combustion gases. Forexample, a burner, as previously discussed, may be used to produce hotcombustion gases (shown as 6507 in FIG. 65B) that are used to heat theworking fluid. An expansion area of cylinder (or warm section) 6522 isdisposed inside the heater head 6506 and defines part of a working gasvolume as discussed above with respect to FIG. 1. A piston 6528 is usedto displace the working fluid contained in the expansion area ofcylinder 6522.

In accordance with an embodiment, crankcase 6502 is welded directly toheater head 6506 at joints 6508 to create a pressure vessel that can bedesigned to hold any pressure without being limited, as are otherdesigns, by the requirements of heat transfer in the cooler. In analternative embodiment, the crankcase 6502 and heater head 6506 areeither brazed or bolted together. The heater head 6506 has a flange orstep 6510 that axially constrains the heater head and transfers theaxial pressure force from the heater head 6506 to the crankcase 6502,thereby relieving the pressure force from the welded or brazed joints6508. Joints 6508 serve to seal the crankcase 6502 (or cold-end pressurevessel) and bear the bending and planar stresses. In an alternativeembodiment, the joints 6508 are mechanical joints with an elastomerseal. In yet another embodiment, step 6510 is replaced with an internalweld in addition to the exterior weld at joints 6508.

Crankcase 6502 is assembled in two pieces, an upper crankcase 6512 and alower crankcase 6516. The heater head 6506 is first joined to the uppercrankcase 6512. Second, a cooler 6520 is installed with a coolant tubing(shown as 6514 in FIG. 65B) passing through holes in the upper crankcase6512. Third, the double acting pistons 6528 and drive components(designated generally as numeral 6540 in FIGS. 65A and 65C, not shown inFIG. 65B) are installed. In one embodiment, lower crankcase 6516 isassembled in three pieces, an upper section 6513, a middle section 6515,and a lower section 6517, as shown in FIGS. 65A and 65C. Middle section6515 is may be connected to upper and lower sections 6513 and 6517 atjoints 6519 and 6521, respectively, by any mechanical means known in theart, or by welding.

The lower crankcase 6516 is then joined to the upper crankcase 6512 atjoints 6518. Preferably, the upper crankcase 6512 and the lowercrankcase 6516 are joined by welding. Alternatively, a bolted flange maybe employed (as shown in FIGS. 65B and 65C).

In some embodiments a motor/generator (shown as 6501 in FIG. 65C), suchas a PM generator, may be installed into motor/generator housing (shownas 6503 in FIG. 65C), which is attached to the lower crankcase 6516, asshown in FIG. 65C. Motor/generator housing 6503 may be attached to lowercrankcase 6516 by any mechanical means known in the art, or may bewelded to lower crankcase 6516. Motor/generator housing 6503 mayassembled in two pieces, a front section 6505, which is attached tolower crankcase 6516, and a rear section 6509, which may be welded orbolted to front section 6505. In one embodiment a seal 6511 may bepositioned between the rear section 6509 and the front section 6505 ofthe motor/generator housing 6503. In some embodiments rear section 6509is removable attached to front section 6505, which serves, among otherfunctions, to allow for easy removal and installation of motor/generator6501 during engine 6500 assembly.

In order to allow direct coupling of the heater head 6506 to the uppercrankcase 6512, the cooling function of the thermal cycle is performedby a cooler 6520 that is disposed within the crankcase 6502, therebyadvantageously reducing the pressure containment requirements placedupon the cooler. By placing the cooler 6520 within crankcase 6502, thepressure across the cooler is limited to the pressure difference betweenthe working gas in the working gas volume, and the charge gas in theinterior volume 6504 of the crankcase. The difference in pressure iscreated by the compression and expansion of the working gas, and istypically limited to a percentage of the operating pressure. In oneembodiment, the pressure difference is limited to less than 30% of theoperating pressure.

Coolant tubing 6514 advantageously has a small diameter relative to thediameter of the cooler 6520. The small diameter of the coolant passages,such as provided by coolant tubing 6514, is key to achieving high heattransfer and supporting large pressure differences. The required wallthickness to withstand or support a given pressure is proportional tothe tube or vessel diameter. The low stress on the tube walls allowsvarious materials to be used for coolant tubing 6514 including, but notlimited to, thin-walled stainless steel tubing or thicker-walled coppertubing.

An additional advantage of locating the cooler 6520 entirely within thecrankcase 6502 (or cold-end pressure vessel) volume is that any leaks ofthe working gas through the cooler 6520 will only result in a reductionof engine performance. In contrast, if the cooler were to interface withthe external ambient environment, a leak of the working gas through thecooler would render the engine useless due to loss of the working gasunless the mean pressure of working gas is maintained by an externalsource. The reduced requirement for a leak-tight cooler allows for theuse of less expensive fabrication techniques including, but not limitedto, powder metal and die casting.

Cooler 6520 is used to transfer thermal energy by conduction from theworking gas and thereby cool the working gas. A coolant, either water oranother fluid, is carried through the crankcase 6502 and the cooler 6520by coolant tubing 6514. The feedthrough of the coolant tubing 6514through upper crankcase 6512 may be sealed by a soldered or brazed jointfor copper tubes, welding, in the case of stainless steel and steeltubing, or as otherwise known in the art.

The charge gas in the interior volume 6504 may also require cooling dueto heating resulting from heat dissipated in the motor/generatorwindings, mechanical friction in the drive, the non-reversiblecompression/expansion of the charge gas, and the blow-by of hot gasesfrom the working gas volume. Cooling the charge gas in the crankcase6502 increases the power and efficiency of the engine as well as thelongevity of bearings used in the engine.

In one embodiment, an additional length of coolant tubing (shown as 6530in FIG. 65B) is disposed inside the crankcase 6502 to absorb heat fromthe charge gas in the interior volume 6504. The additional length ofcoolant tubing 6530 may include a set of extended heat transfer surfaces(shown as 6548 in FIG. 65B), such as fins, to provide additional heattransfer. As shown in FIG. 65B, the additional length of coolant tubing6530 may be attached to the coolant tubing 6514 between the crankcase6502 and the cooler 6520. In an alternative embodiment, the length ofcoolant tubing 6530 may be a separate tube with its own feedthrough ofthe crankcase 6502 that is connected to the cooling loop by hosesoutside of the crankcase 6502.

In another embodiment the extended coolant tubing 6530 may be replacedwith extended surfaces on the exterior surface of the cooler 6520 or thedrive housing (shown as 6572 in FIGS. 65A and 65C). Alternatively, a fan(shown as 6534 in FIG. 65B) may be attached to the engine crankshaft(shown as 6542 in FIG. 65C) to circulate the charge gas in interiorvolume 6504. The fan 6534 may be used separately or in conjunction withthe additional coolant tubing 6530 or the extended surfaces on thecooler 6520 or drive housing 6572 to directly cool the charge gas in theinterior volume 6504.

Preferably, coolant tubing 6514 is a continuous tube throughout theinterior volume 6504 of the crankcase and the cooler 6520.Alternatively, two pieces of tubing could be used between the crankcaseand the feedthrough ports of the cooler. One tube carries coolant fromoutside the crankcase 6502 to the cooler 6520. A second tube returns thecoolant from the cooler 6520 to the exterior of the crankcase 6502. Inanother embodiment, multiple pieces of tubing may be used between thecrankcase 6502 and the cooler in order to add tubing with extended heattransfer surfaces inside the crankcase volume 6504 or to facilitatefabrication. The tubing joints and joints between the tubing and thecooler may be brazed, soldered, welded or mechanical joints.

Various methods may be used to join coolant tubing 6514 to cooler 6520.Any known method for joining the coolant tubing 6514 to the cooler 6520may be used in various embodiments. In one embodiment, the coolanttubing 6514 may be attached to the wall of the cooler 6520 by brazing,soldering or gluing. Cooler 6520 is in the form of a cylinder placedaround the cylinder 6522 and the annular flow path of the working gasoutside of the cylinder 6522. Accordingly, the coolant tubing 6514 maybe wrapped around the interior of the cooler cylinder wall and attachedas mentioned above.

Alternative cooler configurations are presented in FIGS. 65D-65G thatreduce the complexity of the cooler body fabrication. FIG. 65D shows oneembodiment of a side view of a Stirling cycle engine including coolanttubing. In FIG. 65D, cooler 6552 includes a cooler working space 6550.Coolant tubing 6548 is placed within the cooler working space 6550, sothat the working gas can flow over an outside surface of coolant tubing6548. The working gas is confined to flow past the coolant tubing 6548by the cooler body 6552 and a cooler liner 6526. The coolant tube passesinto and out-of the working space 6550 through ports in either thecooler 6552 or the drive housing 6572 (shown in FIGS. 65A and 65C). Thecooler casting process is simplified by having a seal around coolantlines 6548. In addition, placing the coolant line 6548 in the workingspace improves the heat transfer between the working fluid and thecoolant fluid. The coolant tubing 6548 may be smooth or may haveextended heat transfer surfaces or fins on the outside of the tubing toincrease heat transfer between the working gas and the coolant tubing6548. In another embodiment, as shown in FIG. 65E, spacing elements 6554may be added to the cooler working space 6550 to force the working gasto flow closer to the coolant tubes 6548. The spacing elements areseparate from the cooler liner 6526 and the cooler body 6552 to allowinsertion of the coolant tube and spacing elements into the workingspace.

In another embodiment, as shown in FIG. 65F, coolant tubing 6548 isovercast to form an annular heat sink 6556 where the working gas canflow on both sides of the cooler body 6552. The annular heat sink 6556may also include extended heat transfer surfaces on its inner and outersurfaces 6560. The body of the cooler 6552 constrains the working gas toflow past the extended heat exchange surfaces on heat sink 6556. Theheat sink 6556 is typically a simpler part to fabricate than the cooler6520 in FIGS. 65A and 65B. The annular heat sink 6556 provides roughlydouble the heat transfer area of cooler 6520 shown in FIGS. 65A and 65B.In another embodiment, as shown in FIG. 65G, the cooler liner 6526 canbe cast over the coolant lines 6548. The cooler body 6552 constrains theworking gas to flow past the cooler liner 6562. Cooler liner 6526 mayalso include extended heat exchange surfaces on a surface 6560 toincrease heat transfer.

Returning to FIG. 65B, one method for joining coolant tubing 6514 tocooler 6520 is to overcast the cooler around the coolant tubing. Thismethod is described, with reference to FIGS. 66A and 66B, and may beapplied to a pressurized close-cycle machine as well as in otherapplications where it is advantageous to locate a cooler inside thecrankcase.

Referring to FIG. 66A, a heat exchanger, for example, a cooler 6520(shown in FIGS. 65A and 65B) may be fabricated by forming ahigh-temperature metal tubing 6602 into a desired shape. In oneembodiment, the metal tubing 6602 is formed into a coil using copper. Alower temperature (relative to the melting temperature of the tubing)casting process is then used to overcast the tubing 6602 with a highthermal conductivity material to form a gas interface 6604 (and 6532 inFIG. 65B), seals 6606 (and 6524 in FIG. 65B) to the rest of the engineand a structure to mechanically connect the drive housing 6572 (shown inFIG. 2) to the heater head 6506 (shown in FIG. 65B. In one embodiment,the high thermal conductivity material used to overcast the tubing isaluminum. Overcasting the tubing 6602 with a high thermal conductivitymetal assures a good thermal connection between the tubing and the heattransfer surfaces in contact with the working gas. A seal is createdaround the tubing 6602 where the tubing exits the open mold at 6610.This method of fabricating a heat exchanger advantageously providescooling passages in cast metal parts inexpensively.

FIG. 66B is a perspective view of a cooling assembly cast over thecooling coil of FIG. 66A. The casting process can include any of thefollowing: die casting, investment casting, or sand casting. The tubingmaterial is chosen from materials that will not melt or collapse duringthe casting process. Tubing materials include, but are not limited to,copper, stainless steel, nickel, and super-alloys such as Inconel. Thecasting material is chosen among those that melt at a relatively lowtemperature compared to the tubing. Typical casting materials includealuminum and its various alloys, and zinc and its various alloys.

The heat exchanger may also include extended heat transfer surfaces toincrease the interfacial area 6604 (and 6532 shown in FIG. 65B) betweenthe hot working gas and the heat exchanger so as to improve heattransfer between the working gas and the coolant. Extended heat transfersurfaces may be created on the working gas side of the heat exchanger6520 by machining extended surfaces on the inside surface (or gasinterface) 6604. Referring to FIG. 65B, a cooler liner 6526 (shown inFIG. 65B) may be pressed into the heat exchanger to form a gas barrieron the inner diameter of the heat exchanger. The cooler liner 6526directs the flow of the working gas past the inner surface of thecooler.

The extended heat transfer surfaces can be created by any of the methodsknown in the art. In accordance some embodiments, longitudinal grooves6704 are broached into the surface, as shown in detail in FIGS. 67A and67C. Alternatively, lateral grooves 6708 (also shown in enlarged sectionview FIG. 67B-1) may be machined in addition to the longitudinal grooves6704 (also shown in enlarged section view FIG. 67A-1) thereby creatingaligned pins 6710 as shown in FIG. 67B. In some embodiments, grooves arecut at a helical angle to increase the heat exchange area.

In an alternative embodiment, the extended heat transfer surfaces on thegas interface 6604 (as shown in 66B) of the cooler are formed from metalfoam, expanded metal or other materials with high specific surface area.For example, a cylinder of metal foam may be soldered to the insidesurface of the cooler 6604. As discussed above, a cooler liner 6526(shown in FIG. 65B) may be pressed in to form a gas barrier on the innerdiameter of the metal foam. Other methods of forming and attaching heattransfer surfaces to the body of the cooler are described in U.S. Pat.No. 6,694,731, issued Feb. 24, 2004, entitled Stirling Engine ThermalSystem Improvements, which is herein incorporated by reference in itsentirety.

Additional coolant penetrating cold-end pressure vessel embodiments aredescribed in U.S. Pat. No. 7,325,399. It is to be understood that thevarious coolant penetrating cold-end pressure vessel embodimentsreferred to herein may be adapted to function in a multiple cylinderengine configuration.

Intake Manifold

Referring now to FIGS. 68-69B, an intake manifold 6899, is shown forapplication to a Stirling cycle engine or other combustion applicationin accordance with some embodiments. Various embodiments of intakemanifold 6899 are further disclosed in U.S. Pat. No. 6,381,958. Inaccordance with some embodiments, fuel is pre-mixed with air that may beheated above the fuel's auto-ignition temperature and a flame isprevented from forming until the fuel and air are well-mixed. FIG. 68shows one embodiment including an intake manifold 6899 and a combustionchamber 6810. The intake manifold 6899 has an axisymmetrical conduit6801 with an inlet 6803 for receiving air 6800. Air 6800 is pre-heatedto a temperature, typically above 900 K, which may be above theauto-ignition temperature of the fuel. Conduit 6801 conveys air 6800flowing inward radially with respect to combustion axis 6820 to aswirler 6802 disposed within the conduit 6801.

FIG. 69A shows a cross sectional view of the conduit 6801 includingswirler 6802 in accordance with some embodiments. In the embodiment ofFIG. 69A, swirler 6802 has several spiral-shaped vanes 6902 fordirecting the flow of air 6800 radially inward and imparting arotational component on the air. The diameter of the swirler section ofthe conduit decreases from the inlet 6904 to the outlet 6906 of swirler6802 as defined by the length of the swirler section conduit 6801. Thedecrease in diameter of swirler vanes 6902 increases the flow rate ofair 6800 in substantially inverse proportion to the diameter. The flowrate is increased so that it is above the flame speed of the fuel. Atoutlet 6906 of swirler 6802, fuel 6806, which in a one embodiment ispropane, is injected into the inwardly flowing air.

In some embodiments, fuel 6806 is injected by fuel injector 6804 througha series of nozzles 6900 as shown in FIG. 69B. More particularly, FIG.69B shows a cross sectional view of conduit 6801 and includes the fueljet nozzles 6900. Each of the nozzles 6900 is positioned at the exit ofthe swirler vanes 6902 and is centralized between two adjacent vanes.Nozzles 6900 are positioned in this way for increasing the efficiency ofmixing the air and fuel. Nozzles 6900 simultaneously inject the fuel6806 across the air flow 6800. Since the air flow is faster than theflame speed, a flame will not form at that point even though thetemperature of the air and fuel mixture is above the fuel'sauto-ignition temperature. In some embodiments, where propane is used,the preheat temperature, as governed by the temperature of the heaterhead, is approximately 900 K.

Referring again to FIG. 68, the air and fuel, now mixed, referred tohereafter as “air-fuel mixture” 6809, is transitioned in directionthrough a throat 6808 which has a contoured fairing 6822 and is attachedto the outlet 6807 of the conduit 6801. Fuel 6806 is supplied via fuelregulator 6824.

Throat 6808 has an inner radius 6814 and an outer dimension 6816. Thetransition of the air-fuel mixture is from a direction which issubstantially transverse and radially inward with respect to combustionaxis 6820 to a direction which is substantially parallel to thecombustion axis. The contour of the fairing 6822 of throat 6808 has theshape of an inverted bell such that the cross sectional area of throat6808 with respect to the combustion axis remains constant from the inlet6811 of the throat to outlet 6812 of the throat. The contour is smoothwithout steps and maintains the flow speed from the outlet of theswirler to the outlet of the throat 6808 to avoid separation and theresulting recirculation along any of the surfaces. The constant crosssectional area allows the air and fuel to continue to mix withoutdecreasing the flow speed and causing a pressure drop. A smooth andconstant cross section produces an efficient swirler, where swirlerefficiency refers to the fraction of static pressure drop across theswirler that is converted to swirling flow dynamic pressure. Swirlefficiencies of better than 80% may typically be achieved in practice.Thus, the parasitic power drain of the combustion air fan may beminimized.

Outlet 6812 of the throat flares outward allowing the air-fuel mixture6809 to disperse into the chamber 6810 slowing the air-fuel mixture 6809thereby localizing and containing the flame and causing a toroidal flameto form. The rotational momentum generated by the swirler 6802 producesa flame stabilizing ring vortex as well known in the art.

Referring to FIG. 70, a cross-section is shown of combustor 7022 andexhaust gas flow path 7013, as described above in reference to earlierfigures. In accordance with another embodiment it is recognized that thecombustion exhaust gases remain above the temperature of combustion ofthe fuel well beyond the region of combustor 7022, and that, since thefuel/air mixture is typically exceedingly lean, adequate oxidant remainsfor recombustion of the exhaust gases.

FIG. 70 further illustrates the use of a temperature sensor 7002,typically a thermocouple, to monitor the temperature of heater head 7020at the top of external pin array 7030 and thereby to control the fuelflow such as to maintain the temperature at sensor 7002 below atemperature at which the heater head significantly loses strength. Thetemperature at sensor 7002 is preferably maintained approximately50.degree. C. below the melting temperature of the heater head material.

In the configuration depicted in FIG. 70, the use of avariable-cross-section gas flow bypass channel 7004 is illustrated, asdescribed above. The taper of the bypass channel is greatly exaggeratedfor clarity of depiction. Even where a bypass channel is employed, thetemperature profile as a function of distance from the top of the heaterhead is not flat, as would be preferable. Two additional temperaturesensors, 7006 and 7008, are shown at the middle and bottom,respectively, of external pin array 7030, whereby the temperature of theexhaust gas may be monitored.

In accordance some embodiments, additional fuel is added to the exhaustgases at nozzle 7010 via afterburner fuel line 7012. Nozzle 7010 may bea ring burner, circumferentially surrounding heater head 7020 and facingexternal pin array 7030 between the positions designated in FIG. 70 bytemperature sensors 7002 and 7006. The fuel flow through afterburnerfuel line 7012 may be controlled on the basis of the exhaust gastemperature measured by temperature sensor 7008. The precise position oftemperature 7008 is preferably such as to measure the maximumtemperature of the external pin array produced by the combustion of fuelexiting from afterburner nozzle 7010.

Referring to FIG. 71A, a side view is shown in cross section of a burnerand heat recovery system, designated generally by numeral 7100, for athermal cycle engine in accordance some embodiments. In the embodimentshown, heat is exchanged between hot exhaust gases, heated in combustor7022, and air drawn in at air inlet 7104 in a heat exchanger 7106 thatis external to the heater head assembly. Additionally shown is fuelinlet 7108 and igniter 7110 used to initiate ignition in the combustor.Exhaust stream 7112 traverses heat transfer pins 7030 before beingchanneled to heat exchanger 7106. A seal ring 7114 of copper, or othermetal of sufficiently high melting temperature, forms a rod type seal onheater head flange 7116 just below the bottom row of heat transfer pins7030. Copper ring 7114 fits tightly on heater head flange 7116 producinga labyrinth seal. The right-hand portion of the cross-sectional view ofFIG. 71A, showing the region of the seal, is shown, enlarged, in FIG.71B. Copper seal ring 7114 fits tightly on heater head 7101 and has aclose fit within annular groove 7118 on the bottom surface of burnercover 7120. The configuration of ring 7114 in groove 7118 produces alabyrinth seal causing the exhaust gas, in exhaust plenum 7122 to travela convoluted path around the back side of seal ring 7114 therebylimiting exhaust gas leakage. The tight fit of ring 7114 onto head 7101limits exhaust gas leakage axially out of the burner.

It is to be understood that the various intake manifold embodimentsdescribed herein may be adapted to function in a multiple burnerconfiguration.

Gaseous Fuel Burner

Definitions: As used in this section of the detailed description, thefollowing terms shall have the meanings indicated, unless the contextotherwise requires: Fuel-Air Equivalence ratio (.phi.)=Actual Fuel-AirMass Ratio/Stoichiometric Fuel-Air Mass Ratio. The stoichiometricfuel-air mass ratio is defined as the mass ratio needed to balance thefuel+air chemical equation. The stoichiometric fuel-air mass ratio iswell known for common fuels such as propane (0.0638 g fuel/g air) andcalculable for gases such as biogas.

FIG. 72 shows one embodiment of the engine 7212 embodiment having agaseous fuel burner 7201. Various embodiments of the gaseous fuel burner7201 are also disclosed in U.S. patent application Ser. No. 11/122,447,filed May 5, 2005, published Nov. 10, 2005, which is herein incorporatedby reference in its entirety. This embodiment may be used in the contextof a Stirling cycle engine, however, other embodiments of the machineare not limited to such applications. Those skilled in the art willappreciate that the present machine may have application in othersystems, such as, with other types of external combustion engines.

The use of an ejector in a gaseous fuel burner advantageously can solvesome of the challenges faced by the traditional gaseous fuel burners.First, using an ejector can eliminate the need for additional equipment,controls, and space, such as, a gaseous fuel pump, fuel controlcircuitry, and the associated components. Furthermore, using an ejectorsuch as a venturi simplifies the fuel control system by eliminating theneed for a separate fuel control scheme. Based on the corresponding riseof the vacuum with the airflow, and subsequently, an increased fuelflow, the burner power can be regulated by regulating the airflow.Accordingly, removing separate fuel control simplifies the developmentand implementation of automatic burner control in a gaseous fuel burnerwith an ejector.

Secondly, the corresponding rise of the vacuum with airflow also resultsin an approximately steady fuel-air ratio despite changes in temperatureand airflow rates. The resulting steady fuel-air ratio simplifies thefuel control and operation of the burner, by eliminating the need forcomplex exhaust sensor/feedback fuel control mechanisms.

Referring to FIG. 72, a gaseous fuel burner 7201 comprises an ejector7240, a heat exchanger 7220, a combustion chamber 7250, and a blower7200 (shown as 7300 in FIG. 73A). The term ejector as used here includeseductors, siphons, or any device that can use the kinetic energy of onefluid to cause the flow of another fluid. Ejectors are a reliable way ofproducing vacuum-based fuel flow systems with low initial cost, lack ofmoving parts, and simplicity of operation.

Referring again to FIG. 72, in a some embodiments, the ejector 7240 is aventuri. The venturi 7240 is positioned downstream of the outlet of theair preheater or heat exchanger 7220, in a venturi plenum 7241 andproximal to the combustion chamber 7250. A blower 7200 forces airthrough the venturi 7240. The flow of air through the venturi draws in aproportional amount of fuel through the fuel inlet ports 7279. The fuelinlet ports 7279 are placed at the venturi throat 7244 where the throathas the lowest pressure. The ports 7279 are sized to produce plumes offuel across the airflow that promote good mixing within the venturi7240. This fuel-air mixture exits the venturi 7240 and forms aswirl-stabilized flame in the combustion chamber 7250. The venturi 7240draws in an amount of fuel that is substantially linearly proportionalto the airflow regardless of airflow rates and temperature of the airentering the venturi 7240.

In a some embodiments as shown in FIGS. 73A and 73B, placing the venturi7340 between the air preheater 7320 and the combustion chamber 7350promotes a substantially steady air-fuel ratio over a wide range ofairflows and venturi temperatures. FIG. 73A is a schematic drawing ofthe burner including the components of the burner such as a blower 7300,a preheater 7320, a venturi 7340, and fuel supply 7372. The drawing alsoincludes a load heat exchanger or heater head 7390 (also shown in FIGS.76-78 as 7290). The load heat exchanger 7390 is the heat exchanger ofthe engine or process that absorbs the thermal power of the hot gasesleaving the combustion chamber 7350 in the burner at some elevatedtemperature. The partially cooled burned gases then enter the exhaustside of the air preheater, where they are further cooled by incomingcombustion air. FIG. 73B shows the pressure map of the same componentsarranged linearly. The air pressure supplied by the blower, the fuelsupply pressure, and the ambient pressure are all indicated. The massflow rate (m′) of fuel into the burner is controlled by the differencebetween the fuel supply pressure at 7372 and the pressure in the venturithroat 7344 (shown in FIG. 72 as 7244) and the fuel temperature at thedominant restriction:m′.sub.FUEL.varies.(P.sub.FUEL−P.sub.THROAT).sup.0.5/T.sub.FUEL.sup.0.5

The pressure in the throat (P.sub.THROAT) is set by the pressure dropthrough the exhaust side of the preheater 7320 plus the pressure dropthrough the heater head tubes 7390 minus the suction generated by theventuri throat 7344. The pressure drops 7320, 7390 and the throatsuction pressure 7344 are all proportional to the airflow rate and theventuri temperature.P.sub.THROAT.varies.mt.sub.AIR.sup.2*T.sub.VENTURI

Combining these equations shows that the fuel flow will varyapproximately linearly with the airflow:m′.sub.FUEL.varies.[P.sub.FUEL−(m′.sub.AIR.sup.2*T.sub.VENTURI)].sup.0.5/T.sub.FUEL.sup.0.5

Regulating the fuel pressure to near ambient pressure, the fuel flow isapproximately linear with airflow.m′.sub.FUEL.varies.mt.sub.AIR*(T.sub.VENTURI/T.sub.FUEL).sup.0.5Thus, locating the dominant fuel restriction 7378 (shown as 7278 in FIG.72) within the venturi plenum (shown as 7241 in FIG. 72) provides for anapproximately steady fuel-air ratio over a wide range of airflow ratesand venturi temperatures.m′.sub.FUEL/m′.sub.AIR.varies.constant

FIG. 74 shows one embodiment of the ejector such as the venturi. In thisembodiment, the size of the opening of the venturi throat 7244determines the amount of suction present at the throat 7244. In aspecific embodiment, the venturi throat is approximately 0.24 inches indiameter. Referring back to FIGS. 72 and 74, fuel delivery means arecoupled to the venturi 7240. The fuel delivery means may be manifolds,fuel lines or fuel tubes. The fuel delivery means may include othercomponents such as a fuel restriction 7278, fuel inlet ports 7279 andfuel valves (not shown). Fuel supplied by a pressure regulator 7272flows through a manifold 7273 and fuel inlet ports 7279 into therelatively lower pressure in the throat 7244. In one embodiment the fuelinlet ports 7279 provide the largest portion of the pressure drop in thefuel delivery means. Preferably, making the fuel inlet ports the largestrestriction in the fuel delivery means assures that the restrictionoccurs at the venturi temperature and maximizes fuel-air mixing byproducing the largest possible fuel plumes. Referring back to FIG. 72,the fuel and air flow into the divergent cone or diffuser 7248 of theventuri, where static pressure is recovered. In the diffuser 7248, theentrained fuel mixes with the air to form an ignitable fuel air mixturein the combustion chamber 7250. The ignitable fuel-air mixture thenenters the combustion chamber 7250, where the igniter 7260 may ignitethe mixture, and the tangential flow induced by a swirler 7230 creates aswirl-stabilized flame. Using an ejector 7240 to draw the gaseous fuelinto the combustion chamber eliminates the need for a high-pressuregaseous fuel pump to deliver the fuel.

In one embodiment, the venturi 7240 is constructed from high temperaturematerials to withstand high temperatures and maintain its structuralintegrity. For the embodiment of FIG. 74, the dimensions of the venturican be approximately 0.9 inches diameter inlet and outlets with anapproximately 0.24 inches diameter throat. The half angles of theconvergent cone and divergent cones can be 21.degree. and 7.degree.respectively and the throat can be 0.25 inches long. In this embodiment,the venturi can be constructed from Inconel 600. Alternatively, otherhigh temperature metals could be used including, but not limited toStainless Steels 310, 316L, 409 and 439, Hastalloy C76, Hastalloy X,Inconel 625 and other super alloys.

In one embodiment, as shown in FIG. 72, a swirler 7230 is locatedupstream of the venturi 7240 and advantageously creates a tangentialflow of air through the venturi. As is well known in the art, thetangential flow from the swirler can create an annular vortex in thecombustion chamber, which stabilizes the flame. Additionally, theswirler 7230 increases the suction pressure at the venturi throat 7244by increasing the local air velocity over the fuel inlet ports 7279.Adding the swirler allows the venturi throat 7244 to be made larger fora given suction pressure. Furthermore, the swirling action induced bythe swirler 7230 can suppress fluctuations in the combustion chamberpressure from propagating upstream to the venturi 7240. Such pressurefluctuations can temporarily slow or stop the flow of fuel gas into theventuri 7240. The swirler 7230 thereby facilitates a steady fuel-airratio in the combustion chamber for steady airflows. The swirler 7230may be a radial swirler.

In other embodiments, the gaseous burner can be connected to multiplefuel sources. In this configuration, the burner may be fired, lit orignited with a type of fuel and then run with a different type of fuel.The use of multiple fuel sources may require a fuel delivery means tunedfor each fuel. FIGS. 75A, 75B, and 75C show embodiments for two fuelswith significantly different energy densities such propane and naturalgas. In this embodiment, the fuel delivery means for the denser propanemust be approximately three times more restrictive than the fueldelivery means for the less dense natural gas or methane. In theembodiment shown in FIG. 75A, the venturi has different manifolds andfuel ports for each fuel. High-density fuels such as propane wouldrequire the more restrictive fuel inlet ports 7279, while a low-densityfuel such as natural gas would require less restrictive fuel inlet ports7279A. This configuration retains the highest resistance to fuel flow atthe venturi temperature. However, the embodiment of the venturi in FIG.75A may be more difficult to manufacture and have a higher-pressure lossdrop due to the long narrow passage.

Another embodiment for a gaseous burner with multiple fuel sources isshown in FIG. 75B. In this embodiment, a fuel selector valve 7276directs the fuel through an additional fuel restriction such as 7278A or7278B for a dense gas or a less dense gas respectively. The multi-portvalve 7276 allows any number of predefined gases to be burned by thesame burner. Predefined gases such as natural gas, liquid petroleum gas(LPG) or biogas can be burned in the same burner by simply setting aselector valve to the corresponding fuel setting. Alternatively, otherembodiments can have multiple settings for different qualities of biogasas the carbon dioxide fraction in biogas can vary from 50% to 20%. Thefuel restrictors may be placed outside the burner as shown in FIG. 75Bor alternatively they can be located in the entrances to the manifold7273. If restrictions 7278 are placed outside of the burner, thensignificant part of the fuel-delivery-means pressure drop is not at theventuri temperature and thus the fuel-air ratio may vary with theventuri temperature. The burner will run initially leaner and getprogressively richer as the hotter faster air flowing through theventuri exerts a stronger vacuum on the fuel. In addition, moving asignificant part of the pressure drop from the fuel ports 7279, the fuelwill not penetrate as far into the air stream. Nevertheless, locatingmultiple restrictors 7278 for different gases may make the fabricationof the part easier.

An alternative embodiment, that provides significant flexibility in thefuel-air ratio control and fuel gas usages, is shown in FIG. 75C. Inthis embodiment, the two fuel sources, 7272A and 7272B are regulated totheir individual pressure and flows though separate fuel delivery meansadjusted for each fuel. Each fuel delivery means includes two or morerestrictions in parallel 7206A and 7208A, and 7206B and 7208B with oneor more valves 7202A, and 7202B, respectively, to vary the pressure dropof the fuel delivery means. The valves may be manually or automaticallyactuated. Fuel selector 7276 connects fuel delivery means to theventuri, while closing the other fuel off.

The multiple restrictions 7206A and 7208A, and 7206B and 7208B and thevalves 7202A and 7202B allow the pressure drop of the fuel deliverymeans to be adjusted during burner warm-up. Thus the fuel-air ratio canbe roughly maintained as the suction pressure increases with increasingventuri temperature. The multiple restrictions can also adjust forchanging fuel gas density. A changing fuel gas density may occur whenthe gaseous fuel burner is connected to biogas digester, wherein thebiogas digester is the source of fuel. In a biogas digester embodiment,the carbon dioxide (CO.sub.2) content and therefore the energy densitycan vary weekly. In this embodiment, if the CO.sub.2 content increases,the pressure-drop through the fuel delivery means must be reduced toallow higher flows of the less energy dense fuel gas. In addition, themultiple restrictions can improve the ignition of the fuel gas byproviding a richer fuel-air mixture for lighting. The richer mixture isprovided by opening additional valves 7202A or 7202B, which also reducesthe pressure-drop of the fuel delivery means. Once the burner is lit,the valve 7202A or 7202B may be closed to produce a leaner flame. Asdescribed supra, once the burner is lit, the burner may be run on adifferent fuel. A fuel selector may be used to switch the fuel types.Alternatively, an embodiment with a multiple fuel selector facilitatesvarying the fuel-air ratio during the operation of the burner.

Now referring to FIGS. 75B and 75C, the fuel selector 7276 may enablethe burner to be lit on one fuel and run on a different type of fuel.This can be important if one fuel is too weak to ignite, but will burnin a warmed up burner. In one example, the burner may be lit on a higherdensity fuel such as propane. Once the burner is warmed up, the fuelselector 7276 is moved to draw in a low-density biogas.

FIG. 76 depicts an embodiment where an automated controller 7288 adjustsa variable restriction 7292 such as a variable flow valve in the fueldelivery means to hold the exhaust oxygen constant as measured by awide-range lambda sensor or UEGO 7286. In this embodiment, the automatedscheme allows any fuel from biogas to propane to be connected to theburner and the control system can compensate for the changing fueldensity. In this embodiment, the automated controller can restrict thefuel path for dense fuels such as propane and open up the fuel path forlow-density fuels such as methane and biogas. Ignition would beaccomplished by starting the variable restrictor 7292 in the fully openposition, which will produce the richest mixture then closing it untilthe fuel-air mixture is ignited. After ignition, the controller cancontrol the fuel flow to achieve the desired exhaust oxygen level. It isalso envisioned that such an embodiment would allow the fuel air ratioto be adjusted during warm-up to optimize efficiency and burnerstability.

In another embodiment as shown in FIG. 77, the gaseous fuel burner is ahigh efficiency burner for an external combustion engine such as aStirling cycle engine. The burner includes manual controls to controlthe burner. The manual controls include a ball valve 7270 to manuallyselect a fuel type, a trim valve 7274 to adjust the fuel-air ratio and arheostat 7702 to control the blower speed, and subsequently the airflow.The preheated air 7222 in the venturi 7240 draws in the fuel from a fuelsource 7272. The fuel then mixes with the preheated air to create afuel-air mixture. The fuel-air mixture flows into the combustion chamber7250 where it burns. In this embodiment a microprocessor/controller 7288holds the heater head temperature constant as measured by thetemperature sensor 7289 by varying the engine speed. Furthermore, theblower-speed determines the burner power output and thus the enginepower output. In an alternative embodiment, the fuel trim valve 7274 isnot included.

Referring now to FIG. 78 the gaseous fuel burner 7201 is a highefficiency burner for an external combustion engine such as a Stirlingcycle engine. In this embodiment, the burner includes an oxygen sensor7286 located in the exhaust stream 7284 and a microprocessor/controller7288 to automatically restrict the fuel flow with the variablerestrictor 7292. Additionally, the burner includes a blower controller(shown as 7702 in FIG. 77). The blower controller 7702 can be adjustedby the microprocessor/controller 7288 to match the Stirling engine poweroutput with the load. In this embodiment, the burner temperature is heldconstant by varying the engine speed and the engine power output isautomatically adjusted by setting the blower speed. Accordingly, in thisembodiment, the burner can burn most gaseous fuels, including fuelswithout constant properties such as biogas.

In another embodiment as shown in FIG. 79, fuel is delivered directlyinto the venturi at a point proximal to the venturi throat 7244. Thisembodiment includes a swirler 7230 to accommodate the fuel deliverymeans such as a fuel line or fuel tube. The swirler 7230 is preferablyan axial swirler positioned in the venturi 7240 and upstream of theventuri throat 7244. In operation, the delivered fuel is entrained withthe motive air to form the fuel-air mixture. The exemplary manual orautomatic control mechanisms are adaptable to this alternate fueldelivery embodiment.

Referring back to FIG. 74, the gaseous fuel burner further comprises anigniter 7260 and a flame-monitoring device 7210. Preferably, the igniter7260 is an excitable hot surface igniter that may reach temperaturesgreater than 1150.degree. C. Alternatively, the igniter 7260 may be aceramic hot surface igniter or an excitable glow pin.

With continuing reference to FIG. 74, other embodiments include aflame-monitoring device 7210. The flame-monitoring device 7210 providesa signal in the presence of a flame. For the safe operation of the anyburner, it is important that the fuel be shut-off in the event of aflameout. The monitoring device for flame sensing is the flamerectification method using a control circuit and a flame rod.

Flame rectification, well known in the art, is one flame sensingapproach for the small, high efficiency gas burners. The device uses asingle flame rod to detect the flame. The flame rod is relativelysmaller than the grounded heater head and it is positioned within thecombustion flame. In this flame rectification embodiment, the controlunit electronics are manufactured by Kidde-Fenwal, Inc., and the flamerod is commercially available from International Ceramics and HeatingSystems

Preferably, the flame-monitoring device uses the hot surface igniter asthe flame rod. Alternatively, the flame-monitoring device may be eitherremote from the hot surface igniter, or packaged with the igniter as asingle unit.

Alternatively, an optical sensor may be used to detect the presence of aflame. A preferred sensor is an ultraviolet sensor with a clear view ofthe flame brush through an ultraviolet transparent glass and a sighttube.

It is to be understood that the various fuel burner embodimentsdescribed herein may be adapted to function in a multiple burnerconfiguration.

Fuel Pump

In accordance with some embodiments, a fuel flow to a pressurizedcombustion chamber of an engine, such as a Stirling engine, may bemetered by varying the operating parameters of a fuel pump. Variousembodiments of the fuel pump are described below and in U.S. Pat. No.7,111,460, issued Sep. 26, 2006, to Jensen et al., and U.S. patentapplication Ser. No. 11/534,979, filed Sep. 25, 2006, published Feb. 8,2007, which are herein incorporated by reference in their entireties.Desired performance may be achieved without the throttle plates orvalves or other restrictive devices that are normally used to meter thefuel flow to the combustion chamber.

FIG. 80 shows a metering pump system providing gaseous fuel to apressurized combustion chamber 8058 of an engine 8022 according to oneembodiment. A gas train, labeled generally as 8005, includes a fuel pump8014, interconnecting lines 8038, 8042 and may include a pressureregulator 8018. The fuel pump 8014 raises the fuel pressure in line 8038to a higher pressure in line 8042. The gas train delivers fuel from thegas supply to the burner 8010, where it is mixed with air and burned ina combustion chamber 8058. The fuel pump is controlled by a controller8034 that modulates the fuel flow rate by varying one or more parametersof an electrical signal sent to the fuel pump 8014. The controller mayalso regulate a blower 8060 that provides air to the combustion chamber8058 and may receive signals from sensors that report engine-operatingparameters.

In some embodiments the delivered fuel pressure in line 8038 is 6 to 13inches water column for liquefied petroleum gas. Natural gas may besupplied in line 8038 at even lower pressures of 3 to 8 inches watercolumn. Alternatively, pressure regulator 8018 can supply the fuel atlower pressures, even negative pressures. Typical fuel pressures in line8042 may range from 0.5 to 5 PSTG.

In some embodiments, fuel pump 8014 is a linear piston pump. A linearpiston pump is shown in FIG. 81. The pump includes a cylinder 8100, apiston 8102, a winding 8104, a spring 8106 and check valves 8108, 8112.When an electrical signal is applied to winding 8104, the winding pullsthe ferrous metal piston 8102 to the left, compressing the spring 8106.Check valve 8108 in the piston allows fuel to flow into compressionvolume 8110. When the electrical signal is turned off and theelectromagnetic force on the piston begins to decrease, the piston 8102is forced to the right by the spring 8106. Gas is forced out check valve8112 into the receiver volume 8114 at a higher pressure.

The flow rate of the pump can be modulated by varying the stroke of thepiston 8102. In one embodiment the signal from the controller to thepump is a half-wave alternating current (“AC”) signal, as shown in FIG.82. Circuitry to produce this signal is well known in the art. Thepiston stroke and, thus, the flow rate increases as the amplitude of theAC signal increases. In some embodiments, low amplitude signals arebiased slightly higher to improve repeatability and linearity of flowversus the driving signal. The force applied to the piston 8102 by thewindings 8104 is inversely proportional to the distance from thewindings to the piston. At low signal levels, the piston does not getvery close to the windings and small changes in the friction and inertiaof the piston will produce significant changes in the resulting pistonstroke and flow. A bias voltage is applied to bring the resting-positionof the piston closer to the windings, so that small changes in thecontroller signal that drives the piston dominate the frictional forcesand the inertia of the piston. For example, the bias voltage added tothe signal is highest at the lowest driving signal (10% signal in FIG.82) and may drop to zero before the drive signal reaches 50%. The biasis reduced at higher flow levels to take advantage of the full pumpstroke.

In another embodiment, the controller signal that drives the pump is apulse-width-modulated (“PWM”) direct current (“DC”) voltage signal. FIG.83 shows an exemplary DC waveform that may be used to drive the pump.Circuitry to generate the PWM DC signal in FIG. 83 is well known in theart. Three different drive signals are plotted versus time. These signalmodulations correspond to 10%, 50% and 90% duty cycles, which are shownfor purposes of illustration and not for limitation. Applying therectangular wave voltages of FIG. 83 to the windings 8104 of FIG. 81will cause the piston 8102 to move to the left and compress the spring8106. The stroke and, therefore, the flow will be roughly proportionalto the voltage times the duration of the signal. The lower signals, 10%and 50%, include bias voltages between signal pulses. As in the case ofthe AC drive signal, the bias voltage moves the piston closer to thewindings to provide greater piston response to small changes in thesignal and overcome the frictional and inertia forces of the piston.This bias voltage may be varied with the duration of the drive signal.The bias voltage is highest at the minimum drive signal duration and maydrop to zero before the drive voltage pulse duty cycle reaches 50%.

Other embodiments may use different controller signal waveforms to drivethe piston. In another embodiment, the piston pump of FIG. 81 can bedriven without the bias voltages shown in FIGS. 82 and 83.

In another embodiment both the frequency and the duration of the PWM DCcontroller signal modulating the pump can be varied to linearize theflow through the pump with changes in the driving signal.

In further embodiments, pump 8014 is a diaphragm pump as shown in FIG.84. In the diaphragm pump, one or more solenoidal coils 8200 drive theshaft of the pump 8202 back and forth. The shaft 8202 deflects twodiaphragms 8204 that alternatively pull gas into the chambers 8212 andthen expel it. The two wire coil is driven with an AC signal connectedto wires (8234, 8236) that drives the piston 8202 back and forth byreversing the flow of current through the coil 8200. The solenoid has apermanent magnet so that a reversing magnetic field can drive thesolenoid in opposite directions. The pumping force on the two chambers8212 is phased 180 degrees apart so that as one chamber is filled, thecompanion chamber is emptied. Check valves 8208 upstream of the pumpingchambers 8212 allow gas flow in, while the downstream valves 8210 allowflow out of the chambers and into the receiver volume 8216. Thesolenoidal coil 8200 can be driven with a full wave AC signal. Insimilar fashion to the piston pump, varying the amplitude of the ACsignal will vary the stroke and, therefore, the fuel flow through thediaphragm pump.

In another embodiment, the electrical coil 8200 in the diaphragm pump8014 of FIG. 84 can be center-tapped by adding a third wire 8232 to thecenter of the coil 8200. Wires (8234 and 8236) connect to each end ofthe coil. This three wire connection allows the piston 8202 to be drivenback and forth with a DC source. The DC source connects to the centerwire 8232 and the other connecting wires (8234 and 8236) are alternatelyconnected to ground or a negative voltage, causing current to flow inone half-coil or the other.

A three-wire coil 8302 and devices (8304, 8306, 8308) to control the DCcurrent flow to the coil are shown schematically in FIG. 85. The coilmay be used to drive a diaphragm pump solenoid, as in FIG. 85. Devices(8304, 8306, 8308) may be relays, field effect transistors (“FET”),bipolar transistors or other similar devices. The controller can varythe flow of fuel through the diaphragm pump by varying the amplitude ofapplied DC voltage signal 8312 using device 8304. Devices 8306, 8308 canbe driven as shown in FIG. 86A, where first one device is closed, thenopened and then the other device is closed and then opened. The verticalaxis of the figure corresponds to a normalized driving voltage, where asignal equal to “1” means a device is closed (i.e., shorted). Controlstrategies using PWM signals, as illustrated in FIG. 83, albeit withoutthe bias described previously for the piston pump and with suitablephasing, can be applied to each of devices 8306, 8308 in FIG. 85.

In another embodiment the amplitude and frequency of the diaphragm pumpstroke of FIG. 84 can be controlled using the three devices (8302, 8304,8306) shown in FIG. 85. The amplitude of the pump stroke is controlledby the average voltage at wire 8312. This voltage can be modulated byfast pulse-width-modulating device 8304. The stroke frequency may becontrolled as before by devices 8306 and 8308. Alternatively, device8304 can be eliminated and switches 8306 and 8308 can be pulse-widthmodulated at a high frequency during their “on” state, as illustrated inFIG. 86B. In other embodiments the center-tapped coil can be replaced bya full bridge or a half-bridge, as known to those skilled in the art.

In other embodiments for use in applications where a constant flow offuel is important, a filter 8701 may be added between pump 8700 andburner head 8706, where the fuel is mixed with the combustion air, asshown in FIG. 87A. One embodiment of the filter 8701 is an RC filtercomprising a capacitance (volume) 8702 and an orifice 8704. The volumeand orifice are sized to allow the required fuel flow and reducefluctuations in flow to a desired level. Mathematical techniques thatare well known in the art may be used to determine these filterparameters.

An acoustic filter using a volume and an orifice restrictor has theelectrical circuit analog shown in FIG. 87B. The analog of gas flow iselectrical current, the analog of gas pressure is electrical voltage,the analog of volume is electrical capacitance, the analog of flowresistance is electrical resistance and the analog of gas inertia iselectrical inductance. The orifice restrictor does not translatedirectly into this model because the orifice flow resistance isproportional to the gas flow squared (non-linear) instead of beingproportional to the gas flow as the model suggests. The model can beused through the process of linearization of flow resistance for smallsignals. The pump gas flow ripple is attenuated by the factor of1/(1+2.pi.fRC). Where “f” is the frequency component of the gas flowentering the filter from the pump. Due to the orifice restrictornon-linear characteristics, the acoustic filter has a lower attenuationat low flow causing a high burner flow ripple as a percentage of averageflow. The higher ripple can cause flame instability and higher emissionsof pollutants. This non-linearity also causes a high resistance toaverage gas flow at the higher flow rates reducing the pump maximum flowcapability.

The addition of a long thin tube 8703 to the acoustic filter providesripple attenuation through the gas mass acceleration, as shown in FIG.87C. The diagram for the electrical analog is shown in FIG. 87D. Thepump gas flow ripple is attenuated by the factor of1/[1+(LC)(2.pi.f).sup.2]. Since L and C are not a function of flow, thefilter attenuation is not affected by the flow rate and does not havethe disadvantages of the filter of FIG. 87A. Attenuation of the ripplealso increases the pump's flow rate.

Referring again to FIG. 80, in another embodiment, controller 8034modulates the output of the fuel pump 8014 to control the temperature ofthe heater tubes 8026 of the engine. The temperature of the heater tube8026 may be measured with a temperature sensor 8054, such as athermocouple, that is attached to a heater tube 8026. When the engineincreases speed, the engine draws more thermal energy from the heatertubes 8026. The tubes cool and the thermocouple 8054 reports thistemperature drop to the controller 8034, which in turn increases thefuel flow until the measured temperature is restored to a specifiedlevel. Any of the devices and methods for metering the fuel through thefuel pump, as described above, may be employed in this embodiment of themachine. Various fuel pump types including rotary vane pumps,piezoelectric pumps, crank driven piston pumps, etc., may be employed.In other embodiments, various operating parameters of a system, of whichthe pressurized chamber is a part, may be controlled by controlling thefuel pump to meter the fuel flow to the chamber. For example, the speedof an internal combustion engine or the power output of an engine may bedetermined by the controller. Alternatively, a fuel/air mixture ratio toa burner may be maintained by the controller.

It is to be understood that the various fuel pump embodiments describedherein may be adapted to function in a multiple burner configuration.

Single Burner Multiple Piston Engine

Referring now to FIGS. 88, 89A-89C, various embodiments is shown whereinan engine 8800, such as a Stirling cycle engine, having a rocking beamdrive 8802 (also shown as 810 and 812 in FIG. 8) and a plurality ofpistons (also shown in FIG. 8 as 802, 804, 806, and 808), includes asingle burner (shown as 8900 in FIGS. 89A and 89B) to heat heater heads8804 of the pistons. Heater heads 8804 may be one of the variousembodiments disclosed in the preceding section, including, but notlimited to, tube heater heads, as designated by numeral 8902 in FIG. 89A(also shown as 9116 in FIGS. 91C and 91D), or pin or fin heater heads,as designated by numeral 8904 in FIG. 89C (and also shown as 5100 inFIGS. 53D through 53F). FIG. 89B included a pin heater head 8904 havinga heater head lining 8926 fitted around the heater head 8904. Burner8900 may be one of any of the various embodiments disclosed in thepreceding sections and in U.S. Pat. No. 6,971,235, issued Dec. 6, 2005,to Langenfeld et al., which is herein incorporated by reference in itsentirety.

In one embodiment a combustion chamber 8906 is positioned above theheater heads 8900, as shown in FIGS. 89A-89C. A prechamber 8901 mayconnect the combustion chamber 8906 to a burner head 8903 via aprechamber nozzle 8908, wherein prechamber nozzle 502 may be a simplenozzle, a swirler nozzle, or a pressure swirl nozzle. The burner head8903 may house a UV window 8910 for flame detection, a fuel injector8912, which may be an air-assist fuel injector such as a Delevan siphonnozzle, and a hot surface igniter 8914. Also connected to the burnerhead 8903 are a first inlet 8916 and a second inlet 8918. One of theseinlets may be a liquid fuel inlet, and the other inlet may be anatomizing inlet.

The prechamber 8901 is a centrally located fuel preparation stagelocated upstream from the combustion chamber 8906. The prechamber 8901is where the fuel is ignited to form a diffusion flame. In oneembodiment where liquid fuel is used, the liquid fuel passes through thefirst inlet 8916. Atomizer passes through the second inlet 8918 toatomize the liquid fuel and mix with the liquid fuel in the prechamber8901. As the atomizer and liquid fuel enter the prechamber 8901 via fuelinjector 8912, it is ignited by the hot surface igniter 8914. Air mayalso pass through an intake 8920 and be preheated by a preheater 8922before it travels into the prechamber 8901, where it will mix with theatomizer and the liquid fuel. Once the mixture is preheated and formedinto a diffusion flame, it travels through the prechamber nozzle 8908into the combustion chamber 8906 to form a PPV (premixed prevaporized)flame. When the diffusion flame leaves the prechamber 8901, evaporationmay occur in the prechamber 8901 which may allow the diffusion flame tobe relit more easily, should it get flamed out or burned out.

Once the flame is in the combustion chamber 8901, the heat from theflame is used to heat the heater heads 8804. The heated gas from thecombustion chamber 8901 evenly flows over the surface of each of theheater heads 8804, wherein heater heads 8804 transfer the heat containedin the heated gas to a working fluid contained in the working space(shown as 8806 in FIG. 88) of the engine (shown as 8800 in FIG. 88). Thecombustion chamber 8901 may have apertures 8924 in its surface tofurther assist in distributing the PPV flame evenly across each of theheater heads 8804.

As described above in the current and preceding sections, the heaterheads 8804 may be a pin heater head, a folded fin heater head, or may beheater tubes. In an embodiment using a pin or fin heater head, theheater head may include a heater head lining 8926 as shown in FIG. 89B(and also shown as 5340 in FIG. 53A). The heater head lining 8926 may bea sleeve that is fitted around the heater head 8904 or it may be asleeve that is heated and expanded and then fit around the heater headsuch that when the sleeve cools it contracts and creates a snug fitaround the heater head. The heater head lining 8926 ensures uniform flowof the heated gas. Uniform flow prevents uneven temperature distributionaround the heater heads 8804 and ensures thermal efficiency, asdiscussed in detail in the preceding sections. Resultant exhaust fromthe burner may exit the burner through an exhaust 8928.

Because the burner may reach very high temperatures, the metal sued toform the burner may expand. Expansion of certain burner surfaces 8930may interfere with the efficiency of the engine or may damage the heaterheads 8804. In an alternative embodiment a compliant member may bepositioned between the heater heads 8804, or, should it be used, theheater head lining 8926 and the burner surface 8930. The compliantmember acts as a buffer against the expanding metal burner surface 8930so that the burner surface 8930 does not expand into the heater heads8804.

In an alternative embodiment a gaseous fuel, such as propane may beused. In such an embodiment the burner may include a burner head 8903and a combustion chamber 8906. The burner head 8903 may house the UVwindow 8910 for flame detection, a fuel injector 8912, which may be anair-assist fuel injector such as a Delevan siphon nozzle, and a hotsurface igniter 8914. The gaseous fuel may enter the combustion chamber8906 via the fuel injector 8912. Upon exiting the fuel injector 8912,the gaseous fuel would be ignited by the hot surface igniter 8914,thereby creating a flame inside the combustion chamber 8906. Combustionof gaseous fuels is described in detail in the preceding sections.

In yet another embodiment burner 8900 may use both gaseous and liquidfuels. Similar to the exemplary embodiment described earlier, andvarious other embodiments described in preceding sections, the burner8900 would include a combustion chamber 8906, a prechamber 8901, and aburner head 8903. The combustion chamber 8906 may be positioned abovethe heater heads 8804. A prechamber 8901 may connect the combustionchamber 8906 to a burner head 8903 via a prechamber nozzle 8908, whereinprechamber nozzle 8908 may be a simple nozzle, a swirler nozzle, or apressure swirl nozzle. The burner head 8903 may house a UV window 8910for flame detection, a fuel injector 8912, which may be an air-assistfuel injector such as a Delevan siphon nozzle, and a hot surface igniter8914. Also connected to the burner head 8903 are a first inlet 8916 anda second inlet 8918. One of these inlets may be a liquid fuel inlet andthe other inlet may be an atomizing inlet. A switch may be positionedbetween the first inlet 8916 and the second inlet 8918 so that whengaseous fuel is used, the gaseous fuel would flow through the secondinlet 8918, instead of the atomizer as described above. When liquid fuelis used, the switch would be configured such that liquid fuel would flowthrough the first inlet 8916 and atomizer would flow through the secondinlet 8918.

In a further embodiment of the burner, a blower may be coupled to burner8900.

Multiple Burner Multiple Piston Engine

Referring now to FIGS. 90 through 91B, another embodiment is shownwherein each heater head 9002 of engine 9000 may be heated by anindividual burner 9004, as shown in FIG. 90. Heater heads 9002 may beany of the various embodiments described in the preceding sections,including, but not limited to, tube heater heads, as designated bynumeral 9116 in FIGS. 91B-91D, or pin or fin heater heads, as designatedby numeral 9118 in FIG. 91A (and also shown as 5100 in FIGS. 53D through53F). Burner 9004 may be any one of the various embodiments disclosed inthe preceding sections and in U.S. Pat. No. 6,971,235.

Each burner 9004 includes a burner head 9100. Similar to previousdisclosed burner embodiments, the burner head 9100 has an igniter 9101,a fuel injector 9108, and a UV window (shown as 9107 in FIG. 91B) forflame detection. Fuel passes through a first inlet 9106, where it isheated by the igniter 9101 and formed into a flame. Preheated air,heated by the preheater 9102, may be mixed with the fuel in thecombustion chamber 9103. The heated fuel mixture forms a flame insidethe combustion chamber 9103 and heats the heater head 9002. Any exhaustfrom the burner may exit the burner via an exhaust 9105. In analternative embodiment of the burner, an atomizer may be combined withthe fuel via a second inlet 9110. In another embodiment of the burner, ablower may be incorporated to maintain an average air ration amongst theindividual burners 9004.

Yet another embodiment may include a prechamber 9111, as shown in FIG.91B. In this embodiment, the burner may include a combustion chamber9103, a prechamber 9111, and a burner head 9100. Combustion chambers9103 may be positioned above the heater heads 9002. A prechamber 9111may connect the combustion chamber 9103 to a burner head 9100 via aprechamber nozzle 9112, such as a simple nozzle, a swirler nozzle, or apressure swirl nozzle. The burner head 9100 may house the UV window 9107for flame detection, a fuel injection 9108, which may be an air-assistfuel injector such as a Delevan siphon nozzle, and a hot surface igniter9101. Also connected to the burner head 9100 are a first inlet 9106 anda second inlet 9110. One of these inlets may be a liquid fuel inlet andthe other inlet may be an atomizing inlet.

The prechamber 9111 is a centrally located fuel preparation stagelocated upstream from the combustion chamber 9103. The prechamber 9111is where the fuel is ignited to form a diffusion flame. In oneembodiment, where liquid fuel is used, the liquid fuel passes throughthe first inlet 9106. Atomizer passes through the second inlet 9110 toatomize the liquid fuel and mix with the liquid fuel in the prechamber9111. As the atomizer and liquid fuel enter the prechamber 9111 via fuelinjector 9108, it is ignited by the hot surface igniter 9101. Air mayalso pass through an intake and be preheated by a preheater 9102 beforeit travels into the prechamber 9111, where it will mix with the atomizerand the liquid fuel. Once the mixture is preheated and formed into adiffusion flame, it travels through the prechamber nozzle 9112 into thecombustion chamber 9103 to form a PPV (premixed prevaporized) flame.When the diffusion flame leaves the prechamber 9111, evaporation mayoccur in the prechamber 9111 which may allow the diffusion flame to berelit more easily, should it get flamed out or burned out.

Once the flame is in the combustion chamber 9103, the heat from theflame is used to heat the heater heads 9002. The heated gas from thecombustion chamber 9103 evenly flows over the surface of each of theheater heads 9002, wherein heater heads 9002 transfer the heat containedin the heated gas to a working fluid contained in the working space ofthe engine (shown as 9000 in FIG. 90). The combustion chamber 9103 mayhave apertures (shown as 9114 in FIG. 91A) in its surface to furtherassist in distributing the PPV flame evenly across each of the heaterheads 8804.

The principles of the present invention may be applied to all types ofengines, include Stirling engines, and may be applied to other pistonmachines utilizing cylinders such as internal combustion engines,compressors, and refrigerators. However, the present invention may notbe limited to the double-acting four-cylinder Stirling engine.

While the principles of the invention have been described herein, it isto be understood by those skilled in the art that this description ismade only by way of example and not as a limitation as to the scope ofthe invention. Other embodiments are contemplated within the scope ofthe present invention in addition to the exemplary embodiments shown anddescribed herein. Modifications and substitutions by one of ordinaryskill in the art are considered to be within the scope of the presentinvention.

What is claimed is:
 1. A rocking beam drive mechanism for a machinecomprising: a rocking beam having a fixed rocker pivot; a crankshaft; aconnecting rod coupling the crankshaft to the rocking beam, whereby anoscillatory motion of the rocking beam is converted to rotary motion ofthe crankshaft; at least one cylinder; at least one piston, the at leastone piston housed within the at least one cylinder respectively wherebythe at least one piston is capable of substantially linearlyreciprocating within the at least one cylinder; at least one piston rodhaving a proximal end and a distal end, the proximal end being connectedto the at least one piston and the distal end being connected to therocking beam by a coupling assembly, whereby linear motion of the atleast one piston is converted to the oscillatory motion of the rockingbeam; a working space housing the at least one cylinder and the at leastone piston; a crankcase housing the rocking beam, the crankshaft, theconnecting rod, and the coupling assembly and the crankcase being filledwith a crankcase gas; an airlock located between the crankcase and theworking space, the airlock being in fluid communication with the workingspace, the airlock being filled with a working gas, and the airlockhousing a portion of the at least one piston rod between the proximalend and the distal end of the at least one piston rod; a dynamic sealbetween the airlock and the crankcase, the dynamic seal sealablyconnected to the portion of the at least one piston rod, the dynamicseal sealing the working gas in the airlock from the crankcase gas, andthe dynamic seal being a rolling diaphragm comprising non-woven fabric;and an airlock pressure regulator fluidically coupled to the crankcaseand to the airlock, the airlock pressure regulator being configured totransfer the crankcase gas and/or the working gas in the airlock betweenthe crankcase and the airlock, wherein the working gas in the airlock isat a first pressure and the crankcase gas is at a second pressure. 2.The rocking beam drive mechanism of claim 1 wherein the at least onecylinder further comprises a closed end and an open end, the open endfurther comprising a linear bearing connected to the at least onecylinder, the linear bearing having an opening to accommodate thecoupling assembly.
 3. The rocking beam drive mechanism of claim 2, thecoupling assembly comprising at least one link rod, the at least onepiston rod and the at least one link rod coupled together by anothercoupling mechanism, the another coupling mechanism being located beneaththe linear bearing.
 4. The rocking beam drive mechanism of claim 3wherein the another coupling mechanism is a flexible joint.
 5. Therocking beam drive mechanism of claim 3 wherein the another couplingmechanism is a roller bearing.
 6. The rocking beam drive mechanism ofclaim 3 wherein the another coupling mechanism is a hinge.
 7. Therocking beam drive mechanism of claim 3 wherein the another couplingmechanism is a flexure.
 8. The rocking beam drive mechanism of claim 3wherein the another coupling mechanism is a journal bearing joint. 9.The rocking beam drive mechanism of claim 1 wherein the airlock pressureregulator includes a release valve, a pump and an oil filter.
 10. AStirling cycle machine comprising: at least one rocking beam drivemechanism comprising: a rocking beam having a fixed rocker pivot; atleast one cylinder; at least one piston, the at least one piston housedwithin the at least one cylinder respectively whereby the at least onepiston is capable of substantially linearly reciprocating within the atleast one cylinder; at least one piston rod having a proximal end and adistal end, the proximal end being connected to the at least one pistonand the distal end being connected to the rocking beam by a couplingassembly, whereby linear motion of the at least one piston is convertedto oscillatory motion of the rocking beam; a crankcase housing therocking beam, the crankcase being filled with a crankcase gas; aconnecting rod rotatably connected to the rocking beam at a distal endof the connecting rod; a crankshaft rotatably coupled to a proximal endof the connecting rod, whereby the oscillatory motion of the rockingbeam is transferred to the crankshaft; a working space housing the atleast one cylinder and the at least one piston; an airlock locatedbetween the crankcase and the working space, the airlock being in fluidcommunication with the working space, the airlock being filled with aworking gas; a seal sealing the working gas in the airlock from thecrankcase gas, sealably connected to the at least one piston rod,wherein the seal is a rolling diaphragm comprising non-woven fabric; andan airlock pressure regulator fluidically coupled to the crankcase andto the airlock, the airlock pressure regulator being configured totransfer the crankcase gas and/or the working gas in the airlock betweenthe crankcase and the airlock, wherein the working gas in the airlock isat a first pressure and the crankcase gas is at a second pressure. 11.The Stirling cycle machine of claim 10 wherein the at least one cylindercomprises a closed end and an open end, the open end further comprisinga linear bearing connected to the at least one cylinder, the linearbearing having an opening to accommodate the coupling assembly.
 12. TheStirling cycle machine of claim 11, the coupling assembly comprising atleast one link rod, the at least one piston rod and the at least onelink rod coupled together by another coupling mechanism, the anothercoupling mechanism being located beneath the linear bearing.
 13. TheStirling cycle machine of claim 11 further comprising a motor connectedto the crankshaft.
 14. The Stirling cycle machine of claim 11 furthercomprising a generator connected to the crankshaft.
 15. The Stirlingcycle machine of claim 10 further comprising a lubricating fluid pump inthe crankcase.
 16. The Stirling cycle machine of claim 15 wherein thelubricating fluid pump is a mechanical lubricating fluid pump driven bya pump drive assembly, the pump drive assembly being connected to anddriven by the crankshaft.
 17. The Stirling cycle machine of claim 15wherein the lubricating fluid pump is an electric lubricating fluidpump.
 18. The Stirling cycle machine of claim 10 wherein the airlockpressure regulator includes a release valve, a pump and an oil filter.